Control device of internal combustion engine

ABSTRACT

Cylinders of a diesel engine  1  are provided with cylinder pressure sensors  29   a  to  29   d  for detecting combustion chamber pressures. An electronic control unit (ECU)  20  of the engine selects optimum combustion parameters in accordance with a fuel injection mode of fuel injectors  10   a  to  10   d  of the engine and a combustion mode determined by the amount of EGR gas supplied from the EGR valve  35  from among a plurality of types of combustion parameters expressing the combustion state of the engine calculated based on the cylinder pressure sensor output and feedback controls the fuel injection amount and fuel injection timing so that the values of the combustion parameters match target values determined in accordance with the engine operating conditions. Due to this, the engine combustion state is controlled to the optimum state at all times regardless of the fuel injection mode or combustion mode.

TECHNICAL FIELD

The present invention relates to a control system for an internalcombustion engine, more particularly relates to a control system foroptimizing combustion in a diesel engine.

BACKGROUND ART

Due to the recent toughening of exhaust gas controls and demands forreducing noise, there has been a rising demand for optimizing combustionin the combustion chambers of diesel engines as well. To optimizecombustion, accurate control of the fuel injection amount, fuelinjection timing, injection period, etc. becomes necessary even indiesel engines.

However, in a diesel engine, generally the amount of intake air is notadjusted. The engine load is controlled by the fuel injection amount.Therefore, in a diesel engine, combustion is performed in a leanair-fuel ratio region considerably higher than the stoichiometricair-fuel ratio. Further, the air-fuel ratio changes depending on theload. Therefore, in a conventional diesel engine, the air-fuel ratio isnot strictly controlled as in a gasoline engine. In the past, the fuelinjection amount, fuel injection timing, and other fuel injectionparameters also have not been controlled as strictly as with a gasolineengine. Further, in the past, in a diesel engine, the target values ofthe fuel injection amount, injection timing, injection pressure, andother fuel injection parameters have been determined from the engineoperating conditions (speed, accelerator opening degree, etc.) and thefuel injector has been open-loop controlled in accordance with thosetarget values. With open loop control, however, it was impossible toprevent error in the actual fuel injection amount compared with thetarget fuel injection amount and was difficult to accurately control thecombustion state to the targeted state.

Further, to improve the exhaust gas properties and reduce the noise,multi-fuel injection injecting fuel a plurality of times before andafter the main fuel injection so as to optimally adjust the combustionstate is effective. However, for multi-fuel injection, it is necessaryto precisely control the fuel injection amounts and injection timings ofthe plurality of fuel injections.

Further, in a common rail type high pressure fuel injection systemdesigned to be employed in recent diesel engines for improving thecombustion state, since the fuel injection period is short and the fuelinjection pressure changes during the injection, there is the problem ofa susceptibility to error in the fuel injection amount. Therefore, in acommon rail type high pressure fuel injection system, measures have beenadopted such as setting the tolerance of the fuel injectors small so asto improve the fuel injection accuracy, but in practice fuel injectorschange in fuel injection characteristics along with the period of usedue to wear of the parts etc., so with open loop control, it isdifficult to make the fuel injection parameters constantly accuratelymatch with the target values.

In this way, in a diesel engine, error is liable to occur in the fuelinjection amount etc., so even if setting target values giving theoptimal combustion state, in practice sometimes making the fuelinjection amount match with the target value is difficult.

On the other hand, it is known in the art to reduce the combustiontemperature of an engine so as to reduce the NO_(x) and other harmfulemissions by recirculating part of the engine exhaust gas to the enginecombustion chambers as EGR gas. Further, even in a diesel engine engagedin normal lean air-fuel ratio operation, it is possible to supply arelatively large amounts of EGR gas to the combustion chamber so as toreduce the harmful emissions in the exhaust gas.

However, EGR gas has a large effect on combustion. In particular, in adiesel engine, the amount of EGR gas has a large effect on the ignitiondelay time from the start of fuel injection to when the injected fuelstarts to burn. Therefore, if EGR gas is excessively supplied to thecombustion chamber, the engine combustion state will deteriorate and adrop in the engine performance and deterioration of the exhaust gasproperties will occur.

On the other hand, if the amount of EGR gas is small, the effect ofsuppression of the harmful emissions will fall. Therefore, the amount ofEGR gas has to be controlled to a suitable amount in accordance with theoperating conditions of the engine.

However, conventionally the amount of EGR gas has not been controlledprecisely. In particular, in a diesel engine, the opening degree of anEGR valve for controlling the flow rate of EGR gas usually was open loopcontrolled to a value determined from the engine speed and acceleratoropening degree (amount of depression of accelerator pedal).

However, due to the recent toughening of exhaust gas controls anddemands for reducing noise, a need has arisen to precisely control eventhe EGR gas flow rate to its optimal value. If precisely controlling theEGR in this way, a sufficient accuracy cannot be obtained with open loopcontrol based on the engine speed and accelerator opening degree like inthe past. Further, for example, it is possible to arrange an air-fuelratio sensor in the engine exhaust passage and to control the amount ofEGR gas based on the exhaust air-fuel ratio detected by the air-fuelratio sensor, but with an engine like a diesel engine which is operatedin a state where the exhaust air-fuel ratio is extremely lean, thedetection accuracy of the air-fuel ratio sensor falls, so there is theproblem that if controlling the amount of EGR gas based on the exhaustair-fuel ratio detected by the air-fuel ratio sensor, the error becomeslarge.

Even in control of the amount of EGR gas, it is effective to detect theactual combustion state by some form or another and feedback control theamount of EGR gas so that the actual combustion state matches with thetargeted combustion state.

That is, while conventionally the fuel injection amount, fuel injectiontiming, amount of EGR gas, etc. have been open loop controlled based onthe engine speed and accelerator opening degree, to improve theproperties of the exhaust gas and reduce the noise, it becomes necessaryto accurately feedback control the fuel injection amount, fuel injectiontiming, amount of EGR gas, etc. based on the actual combustion state.

Control of the fuel injection or EGR based on the actual enginecombustion state is for example described in Japanese Unexamined PatentPublication (Kokai) No. 2000-54889.

The system of Japanese Unexamined Patent Publication (Kokai) No.2000-54889 does not relate to a diesel engine, but relates to a gasolineengine, but uses the heat release (heat generation) rate in thecombustion chamber as a combustion parameter expressing the combustionstate of the engine and controls the flow rate of EGR gas and the fuelinjection timing, fuel injection amount, ignition timing, etc. so thatthe heat release rate becomes a predetermined pattern.

That is, the system of Japanese Unexamined Patent Publication (Kokai)No. 2000-54889 arranges a cylinder pressure sensor for detecting thepressure inside an engine combustion chamber in each cylinder,calculates the heat release rate at each crank angle based on thedetected actual pressure inside the combustion chamber (combustionpressure) and crank angle, and feedback controls the amount of EGR gas,ignition timing, fuel injection timing, etc. so that the pattern ofchange of the heat release rate with respect to the crank angle matchesa predetermined ideal pattern of change in accordance with the operatingconditions and thereby obtain the optimal combustion.

The system of Japanese Unexamined Patent Publication (Kokai) No.2000-54889 takes note of the heat release rate as a parameter relatingto combustion, calculates the pattern of the heat release rate in actualoperating conditions, and makes the heat release rate follow apredetermined pattern by feedback controlling the ignition timing, fuelinjection amount, etc. The system of Japanese Unexamined PatentPublication (Kokai) No. 2000-54889 relates to a gasoline engine, but itmay be considered to similarly provide cylinder pressure sensors in adiesel engine as well and thereby calculate the pattern of the heatrelease rate based on the outputs of the cylinder pressure sensors andfeedback control the fuel injection timing and fuel injection amount sothat the peak positions or pattern of the heat release rate becomespredetermined peak positions or pattern of heat release rate.

However, the system of Japanese Unexamined Patent Publication (Kokai)No. 2000-54889 uses only the heat release rate in a combustion chamberas a parameter expressing the combustion state of the engine forfeedback control of the combustion state. The system of JapaneseUnexamined Patent Publication (Kokai) No. 2000-54889 is used by agasoline engine. In a gasoline engine, the pre-mixed air-fuel mixture isignited by sparks. The ignition, combustion, and other combustionparameters also do not change much. Therefore, no great error occurseven if using only the peak positions or pattern of the heat releaserate as a parameter expressing the combustion state.

In a diesel engine however, for example, sometimes not only main fuelinjection, but also multi-fuel injection including pilot injectionperformed before the main fuel injection, after injection performedafter the main fuel injection, etc. is performed. Even looking at justthe type of injection (injection mode), sometimes there is a largedifference. Further, in a diesel engine, the combustion pattern(combustion mode) changes depending on the amount of EGR gas.

For this reason, since the change in pressure in a combustion chambergreatly differs depending on the injection mode or combustion mode aswell, feedback control of the combustion state by just the peakpositions or pattern of the heat release rate is not necessarilysuitable.

For example, in an in-cylinder fuel injector of a diesel engine, theinjection amount, injection timing, and other fuel injectioncharacteristics gradually change along with the period of use resultingin deviation in fuel injection characteristics, but such deviation infuel injection characteristics is difficult to accurately correct basedon the peak positions or pattern of the heat release rate.

Further, when performing pilot injection or main fuel injection or afterinjection or other multi-fuel injection, optimization of the combustionstate requires optimal control of the fuel injection amount andinjection timing of the fuel of each, but feedback control of the fuelinjection characteristics of a plurality of fuel injections is difficultbased on only the peak positions or pattern of the heat release rate.

On the other hand, as an example of a combustion control system of aninternal combustion engine using a parameter other than the heat releaserate to detect the combustion state and controlling the fuel injectioncharacteristics of multi-fuel injection in accordance with thecombustion state, there is the one described in Japanese UnexaminedPatent Publication (Kokai) No. 2001-123871.

The system of Japanese Unexamined Patent Publication (Kokai) No.2001-123871 measures the combustion noise of a diesel engine, judgeswhether the pilot injection amount is too great based on the measuredcombustion noise, and corrects the pilot injection amount based on this.Further, as the combustion noise, it uses the derivative or secondderivative of the cylinder pressure detected by a cylinder pressuresensor detecting the pressure inside a combustion chamber so as toremove the effect of mechanical vibration and thereby improve thedetection accuracy of the combustion noise.

That is, the system of Japanese Unexamined Patent Publication (Kokai)No. 2001-123871 feedback controls the pilot injection amount based onthe actually measured combustion noise so as to keep the combustionnoise below a target level at all times.

As explained above, since the system of Japanese unexamined PatentPublication (Kokai) No. 2001-123871 feedback controls the pilotinjection amount based on the actually measured combustion noise, it cankeep the combustion noise below a target level at all times. However,while the system of Japanese Unexamined Patent Publication (Kokai) No.2001-123871 keeps the combustion noise below a target value, it does notnecessarily always obtain a good combustion state. Conversely, sometimesit deteriorates the exhaust properties.

That is, to obtain good exhaust properties, it is necessary to suitablycontrol not only the injection amount of the pilot injection, but alsothe injection timing, but the system of Japanese Unexamined PatentPublication (Kokai) No. 2001-123871 controls only the injection amountof the pilot injection based on the combustion noise and does notcontrol the injection timing based on the actual combustion state.Therefore, the system of Japanese Unexamined Patent Publication (Kokai)No. 2001-123871 has the problem that while the combustion noise falls,the exhaust properties are not always improved.

Further, the system of Japanese Unexamined Patent Publication (Kokai)No. 2001-123871 deals only with pilot injection and even more so onlyoperation with just one pilot injection, so has the problem that itcannot suitably control the injection amounts and injection timings ofthe different fuel injections in multi-fuel injection consisting of aplurality of pilot injections or after injection performed after mainfuel injection.

DISCLOSURE OF THE INVENTION

The present invention, in view of the above problems, has as its objectto provide a control system for an internal combustion engine using anoptimal combustion parameter in accordance with an injection mode orcombustion mode for feedback control of the fuel injection amount,injection timing, and amount of EGR gas even in a diesel engine so as toenable optimal control of the combustion state of the diesel engine.

To achieve the above object, according to the present invention, thereis provided a control system for an internal combustion engine providedwith a fuel injector for injecting fuel into an engine combustionchamber, an EGR system for recirculating part of the engine exhaust intothe engine combustion chamber as EGR gas, and a cylinder pressure sensorfor detecting a pressure inside the engine combustion chamber, thecontrol system for an internal combustion engine provided withcombustion parameter calculating means for calculating a combustionparameter expressing an engine combustion state including at least oneof a cylinder heat release amount, a combustion start timing, and acombustion period based on a relationship predetermined using thecombustion chamber pressure detected by the cylinder pressure sensor andan engine crank angle and correcting means for correcting at least oneof a fuel injection amount, fuel injection timing, and amount of EGR gasso that the calculated combustion parameter becomes a target valuepredetermined in accordance with the engine operating conditions, acombustion parameter selected in accordance with a fuel injection modeor combustion mode of the engine among a plurality of types ofcombustion parameters expressing the engine combustion state calculatedbased on the combustion chamber pressure and engine crank angle beingused as the combustion parameter for correction by the correcting means.

That is, in the present invention, the combustion parameter expressingthe engine combustion state is calculated based on the actual combustionchamber pressure detected by a cylinder pressure sensor and crank angle,but for example only the heat release rate is not used as the combustionparameter for controlling all cases. The optimal combustion parameterfor the combustion mode determined by the number of fuel injections andother fuel injection modes and the amount of EGR etc., that is, theparameter with the least error in the fuel injection mode or combustionmode, is selected from among a plurality of types of combustionparameters calculated based on the combustion chamber pressure and crankangle and used for the feedback control. By selecting the combustionparameter giving the smallest error in accordance with the fuelinjection mode or combustion mode from among the plurality of types ofcombustion parameters in this way, it becomes possible to optimallycontrol the combustion of a diesel engine.

Note that in the present specification, a parameter expressing thecombustion state in a combustion chamber calculated based on thecombustion chamber pressure is called a “combustion parameter”.

As the combustion parameters able to be used in the present invention,for example, there are the following:

Maximum value Pmax of combustion chamber pressure after start ofcombustion (see FIG. 2) and crank angle where combustion chamberpressure becomes the maximum.

Crank angle when maximum value (dP/dθ)max (see FIG. 4) of rate of changeof combustion chamber pressure to crank angle occurs.

Maximum value PVmax of product of combustion chamber pressure andcombustion chamber actual volume and crank angle where PVmax occurs.

Difference ΔPVmax (=PVmax−PVmaxbase) between maximum value PVmax ofproduct of combustion chamber pressure and combustion chamber volume andproduct PVmaxbase of combustion chamber pressure due only to compressionin case of assuming that no combustion has occurred and combustionchamber actual volume at crank angle (θpvmax) where PVmax occurs (seeFIG. 7).

Crank angle where maximum value (dQ/dθ) max of cylinder heat releaserate occurs.

Overall amount of cylinder heat release ΣdQ.

Difference Pmax−Pmin between maximum value Pmax of cylinder pressureafter start of combustion and cylinder minimum pressure Pmin (see FIG.8) in interval until combustion is started in combustion chamber aftercompression top dead center.

Difference Pmax−Pmaxbase between maximum value Pmax of cylinder pressureafter start of combustion and combustion chamber pressure Pmaxbase (seeFIG. 9) due only to compression in case of assuming no combustionoccurred at crank angle giving maximum cylinder pressure.

By using the suitable parameter for the injection mode or combustionmode from among these combustion parameters so as to control the fuelinjection amount, injection timing, amount of EGR gas, etc., thecombustion state of the engine is optimally controlled.

Further, in particular for control of the EGR, if using as thecombustion parameter the time Δt from the start of fuel injection untilthe crank angle where the maximum value PVmax of the product of theabove-mentioned combustion chamber pressure and combustion chamberactual volume occurs, good precision EGR control becomes possible.

Similarly, for control of the EGR, even if using as the combustionparameter the time Δtc after the start of fuel injection until the valueof PV^(κ) calculated based on the combustion chamber pressure P, thecombustion chamber volume V determined from the crank angle θ, and thespecific heat ratio κ of the combustion gas becomes the minimum valuePV^(κ)min or the time Δtc after the start of fuel injection from a fuelinjector when the value of PV^(κ) becomes the minimum value PV^(κ)min towhen it becomes the maximum value PV^(κ)max, good precision EGR controlbecomes possible.

Further, when performing multi-fuel injection, it is also possible touse as a combustion parameter the combustion period including thecombustion start timing and end timing in a combustion chambercalculated using the rate of change d(PV^(γ))/dθ of the parameter PV^(γ)expressed as a product of the γ power of V and P with respect to thecrank angle θ using the combustion chamber pressure P, the combustionchamber volume V determined from the crank angle θ, and a predeterminedconstant γ and to correct the injection timings and injection amounts(fuel injection pressures) of the different fuel injections so thatthese combustion parameters match with target values.

BRIEF DESCRIPTION OF DRAWINGS

FIG. 1 is a view of the schematic configuration of an embodiment in thecase of applying the fuel injection system of the present invention to avehicular diesel engine,

FIG. 2 is a view for explaining a combustion parameter Pmax,

FIG. 3 is a view for explaining selective use of a combustion parameter(dP/dθ)max in accordance with the injection mode,

FIG. 4 is a view for explaining a combustion parameter (dP/dθ)max,

FIG. 5 is a view explaining selective use of a combustion parameter(d2P)/dθ2)max in accordance with the injection mode,

FIG. 6 is a view explaining selective use of a combustion parameter(dQ/dθ)max in accordance with the injection mode,

FIG. 7 is a view explaining a combustion parameter (ΔPVmax),

FIG. 8 is a view explaining a combustion parameter (Pmax−Pmin),

FIG. 9 is a view explaining a combustion parameter (Pmax−Pmaxbase),

FIG. 10 is a view explaining a combustion parameter (PVmain−PVmainbase),

FIG. 11 is a view explaining a combustion parameter (Pmtdc−Pmin),

FIG. 12 is a flow chart explaining an embodiment of a fuel injectioncorrection operation of the present invention,

FIG. 13 is a view explaining a principle of calibration of a cylinderpressure sensor,

FIG. 14 is a view explaining a combustion parameter (ΔPVmax−ΔPVafter),

FIG. 15 is a view explaining a combustion parameter (Pmain−Pmainbase),and

FIG. 16 is a flow chart explaining a fuel injection control operation atthe time of switching combustion modes.

Further, FIG. 17 is a view explaining the definitions of combustionparameters used in the present embodiment,

FIG. 18 is a flow chart explaining basic control of the fuel injectionetc. in the present embodiment,

FIG. 19 is a flow chart explaining the control operation of fuelinjection etc. using a combustion parameter in the present embodiment,

FIG. 20 is a flow chart explaining another embodiment of EGR ratecontrol using a combustion parameter,

FIG. 21 is a timing chart explaining control for switching from a normalcombustion mode to a low temperature combustion mode, and

FIG. 22 is a timing chart explaining control fox switching at the timeof reset from the low temperature combustion mode to the normalcombustion mode.

Further, FIG. 23 is a view explaining the different fuel injectionsforming multi-fuel injection,

FIG. 24(A) is a view explaining the principle of detection of thecombustion period in an embodiment,

FIG. 24(B) is a view explaining the principle of detection of an amountof heat release,

FIG. 25 is a flow chart explaining an operation for calculation of thecombustion period and amount of heat release in the different fuelinjections, and

FIG. 26 is a flow chart explaining the routine of a fuel injectioncorrection operation of the present embodiment.

BEST MODE FOR CARRYING OUT THE INVENTION

Below, embodiments of the present invention will be explained using theattached drawings.

FIG. 1 is a view of the schematic configuration of an embodiment in thecase of applying the fuel injection system of the present invention to avehicular diesel engine.

In FIG. 1, 1 indicates an internal combustion engine (in the presentembodiment, a four-cylinder four-cycle diesel engine provided with a #1to #4, that is, four, cylinders being used), while 10 a to 10 d indicatefuel injectors injecting fuel directly into the combustion chambers ofthe #1 to #4 cylinders of the engine. The fuel injectors 10 a to 10 dare connected through fuel passages (high pressure fuel pipes) to acommon rail 3. The common rail 3 has the function of storing thepressurized fuel supplied from a high pressure fuel injection pump 5 anddistributing the stored high pressure fuel through the high pressurefuel pipes to the fuel injectors 10 a to 10 d.

This embodiment is provided with an EGR system for recirculating part ofthe exhaust gas of the engine to the combustion chambers of thecylinders of the engine. The EGR system is provided with an EGR passage33 for connecting the exhaust passage of the engine with the intakepassage of the engine or the intake ports of the cylinders and an EGRvalve 35 arranged in the EGR passage and having the function of a flowcontrol valve for controlling the flow rate of the exhaust gas (EGR gas)recirculated from the exhaust passage to the intake passage. The EGRvalve 35 is provided with a suitable type of actuator such as a steppermotor. The EGR valve opening degree is controlled in accordance with acontrol signal from a later explained ECU 20.

In FIG. 1, 20 shows an electronic control unit (ECU) for controlling theengine. The ECU 20 is configured as a microcomputer of a knownconfiguration including a read only memory (ROM), a random access memory(RAM), a microprocessor (CPU), and input/output ports connected by atwo-way bus. The ECU 20, in the present embodiment, controls the amountof discharge of the fuel pump 5 to control the pressure of the commonrail 3 to a target value determined in accordance with the engineoperating conditions so as to perform fuel pressure control and alsosets the injection timing and injection amount of the fuel injection andthe amount of EGR gas in accordance with the engine operating conditionsand feedback controls the fuel injection amount, injection timing,amount of EGR gas, etc. so that the value of the combustion parametercalculated based on the output of the later explained cylinder pressuresensor matches the target value determined in accordance with the engineoperating conditions so as to perform basic control of the engine.

For this control, in the present embodiment, the common rail 3 isprovided with a fuel pressure sensor 27 for detecting the fuel pressureinside the common rail, while the accelerator pedal (not shown) of theengine 1 is provided near it with an accelerator opening degree sensor21 for detecting the accelerator opening degree (amount of depression ofaccelerator pedal by driver). Further, in FIG. 1, 23 shows a cam anglesensor for detecting a rotational phase of a camshaft of the engine 1,while 25 shows a crank angle sensor for detecting a rotational phase ofa crankshaft. The cam angle sensor 23 is arranged near the camshaft ofthe engine 1 and outputs a reference pulse every 720 degrees convertedto crank rotational angle. Further, the crank angle sensor 25 isarranged near the crankshaft of the engine 1 and generates a crank anglepulse every predetermined crank rotational angle (for example, every 15degrees).

The ECU 20 calculates the engine speed from the frequency of the crankrotational angle pulse signal input from the crank angle sensor 25 andcalculates the fuel injection timings and fuel injection amounts fromthe fuel injectors 10 a to 10 d and the opening degree of the EGR valve35 (amount of EGR gas) based on the accelerator opening degree signalinput from the accelerator opening degree sensor 21 and the enginespeed.

Further, in FIG. 1, 29 a to 29 d show known types of cylinder pressuresensors arranged at the cylinders 10 a to 10 d and detecting thepressures in the cylinder is combustion chambers. The combustion chamberpressures detected by the cylinder pressure sensors 29 a to 29 d aresupplied through an AD converter 30 to the ECU 20.

In the present embodiment, the fuel pressure of the common rail 3 iscontrolled by the ECU 20 to a pressure in accordance with the engineoperating conditions. For example, it is a high pressure of 10 MPa to150 MPa or so and changes in a broad range. Further, with a dieselengine, generally before main fuel injection, pilot injection isperformed for injecting a relatively small amount of fuel into thecylinder once or a plurality of times. The fuel injected into thecylinder by pilot injection burns before main fuel injection and raisesthe temperature and pressure inside the cylinder to a state suited tothe combustion of main fuel injection, so pilot injection enables thecombustion noise to be reduced.

Further, in a diesel engine injecting high pressure fuel like in thepresent embodiment, sometimes after injection or post injection isperformed one or more times after main fuel injection. After injectionis performed when the fuel injection amount of main fuel injectionbecomes great and injection at one time would cause the combustion stateto deteriorate or to optimize the change in combustion pressure in thecylinder, while post injection is performed for raising the exhausttemperature for example.

In a diesel engine, there is no need to accurately control the air-fuelratio to the extent of a conventional gasoline engine, so not muchprecision was demanded in control of the fuel injection amount either.With high pressure fuel injection like the above, however, not only mainfuel injection, but also multi-fuel injection including pilot injection,after injection, etc. is required. (In the present description, thepilot injection, after injection, post injection, etc. performed otherthan the main fuel injection are referred to all together as “multi-fuelinjection”.)

Note that multi-fuel injection will be explained later in detail.

Therefore, in a diesel engine, the fuel injection has to be performedwith a high precision. However, conventional fuel injection control isbasically open loop control determining the fuel injection amount andfuel injection timing from maps preset based on the engine operatingconditions (speed and accelerator opening degree). In actuality, thereare factors causing error in the fuel injection amount such asshortening of the fuel injection time accompanying the increase in fuelinjection pressure, the fluctuation in common rail pressure (fuelinjection pressure) during fuel injection in common rail type fuelinjection, changes in the fuel injection characteristics of a fuelinjector accompanying use, etc. This makes accurate open loop control ofthe fuel injection amount, injection timing, etc. difficult.

Consequently, in the present embodiment, a parameter expressing thecombustion state of the engine is used and the fuel injection amount,injection timing, etc. are feedback controlled so that the parameterbecomes the optimal value set in accordance with the engine operatingconditions (target value) to thereby maintain the combustion state ofthe engine at the optimal state.

The present embodiment uses as the parameter expressing the combustionstate a parameter calculated based on the combustion chamber pressuresdetected by the cylinder pressure sensors 29 a to 29 d and the crankangle. A parameter expressing combustion in a combustion chambercalculated based on the combustion chamber pressure and crank angle isreferred to as a “combustion parameter”.

Note that there are countless parameters expressing the combustion statecalculated based on the combustion chamber pressure, that is, combustionparameters. Theoretically, any of these may be used for feedback controlof the fuel injection amount, injection timing, etc. In actualityhowever, it is learned that the precision of feedback control changesgreatly in some cases depending on the combustion parameter used due tothe fuel injection mode of the engine (main fuel injection alone or mainfuel injection and multi-fuel injection in combination) and combustionmode (magnitude of amount of EGR etc.).

Therefore, the present embodiment sets a plurality of types ofcombustion parameters exhibiting good correlation with the combustionstate in advance and selectively uses from among them the one giving theleast control error in accordance with the fuel injection mode orcombustion mode of the engine 1.

The same combustion parameter is never used regardless of the fuelinjection mode or combustion mode. As explained above, by selectivelyusing the optimal combustion parameter in accordance with the fuelinjection mode or combustion mode from the plurality of types ofcombustion parameters, the present embodiment can maintain the optimalcombustion state of a diesel engine at all times regardless ofdifferences in the fuel injection mode or combustion mode.

Below, typical examples of the combustion parameters used in the presentembodiment will be shown. In the present embodiment, the optimalcombustion parameter for the fuel injection mode or combustion mode isselected for use as the combustion parameter from the followingcombustion parameters:

-   -   (1) Maximum value Pmax of combustion chamber pressure after        start of combustion and crank angle θpmax where this maximum        value occurs (FIG. 2)    -   (2) Crank angle when local maximum value (maximal value)        (dP/dθ)max of rate of change of combustion chamber pressure to        crank angle occurs    -   (3) Crank angle where local maximum value (maximal value)        (d2P/dθ2)max of second derivative of combustion chamber pressure        occurs    -   (4) Local maximum value PVmax of product of combustion chamber        pressure and combustion chamber actual volume and crank angle        θpvmax where that maximum value occurs    -   (5) Difference ΔPVmax (=PVmax−Pvmaxbase) between the above PVmax        and product PVmaxbase between combustion chamber pressure due        only to compression in case of assuming that no combustion has        occurred and combustion chamber actual volume at crank angle        θpvmax where PVmax occurs (see FIG. 7)    -   (6) Crank angle where maximum value (dQ/dθ) max of cylinder heat        release rate occurs    -   (7) Overall amount of cylinder heat release ΣdQ    -   (8) Difference Pmax−Pmin between maximum value Pmax of cylinder        pressure after start of combustion and cylinder minimum pressure        Pmin in interval after compression top dead center until        combustion is started in combustion chamber    -   (9) Difference Pmax−Pmaxbase between maximum value Pmax of        cylinder pressure after start of combustion and combustion        chamber pressure (motoring pressure) Pmaxbase due only to        compression in case of assuming no combustion occurred at crank        angle where Pmax occurs    -   (10) Difference (PVmain−Pmainbase) between product PVmain of        combustion chamber pressure when fuel injected by main fuel        injection is ignited and combustion chamber actual volume and        product PVmainbase of combustion chamber pressure due only to        compression in case of assuming no combustion occurred and        combustion chamber actual volume at crank angle where fuel        injected by main fuel injection is ignited (see FIG. 10)    -   (11) Difference (ΔPVmax−ΔPVafter) between the above PVmax and        the difference ΔPVafter between the product of the combustion        chamber pressure when the fuel injected by after injection is        ignited and the combustion chamber actual volume and the product        of the combustion chamber pressure due only to compression in        case of assuming no combustion occurred and combustion chamber        actual volume at crank angle where fuel injected by after        injection is ignited (see FIG. 10 and FIG. 14)    -   (12) Difference Ptdc−Pmin of combustion chamber pressure at top        dead center of compression stroke and combustion chamber minimum        pressure Pmin in interval after compression top dead center to        when combustion is started in combustion chamber

Below, these combustion parameters will be simply explained.

(1) Pmax and Crank Angle (θpmax) where Pmax Occurs

The maximum value of the combustion chamber pressure after the start ofcombustion usually appears after top dead center of the cylindercompression stroke and is expressed as the cylinder pressure when thecombustion of the fuel injected by the main fuel injection causes thecylinder pressure to rise the most.

FIG. 2 shows the change in cylinder pressure in the expansion strokefrom the suction stroke of a general diesel engine, wherein the ordinateindicates the pressure and the abscissa indicates the crank angle.

In FIG. 2, TDC indicates top dead center in the compression stroke(hereinafter referred to simply as “top dead center”). In a dieselengine, fuel is usually injected near right before top dead center.Combustion is started after the piston passes top dead center, so thecylinder pressure greatly rises after top dead center. Pmax, as shown inFIG. 2, is the maximum value of the combustion chamber pressure afterthe start of combustion, while θpmax is the crank angle when Pmaxoccurs.

Pmax and θpmax are easily found by monitoring the outputs of thecylinder pressure sensors 29 a to 29 d.

In the present embodiment, Pmax is used for correction of the injectionamount of the main fuel injection, while θpmax is used for correction ofthe injection timing of the main fuel injection.

That is, in the present embodiment, the values (target values) of Pmaxand θpmax in the optimal combustion state in the case of operating anengine while changing the engine operating conditions, that is, theengine speed and accelerator opening degree, are found in advance byexperiments and are stored in advance in the ROM of the ECU 20 in theform of two-dimensional maps using the engine speed and the acceleratoropening degree as parameters.

During engine operation, the ECU 20 finds the Pmax and θpmax of thecylinders based on the outputs of the cylinder pressure sensors 29 a to29 d and reads out the target values of Pmax and θPmax from the enginespeed and accelerator opening degree at that time using the above maps.Further, the fuel injection amount of the main fuel injection iscorrected to increase or decrease it so that Pmax matches its targetvalue, while the fuel injection timing of the main fuel injection iscorrected so that θpmax matches its target value.

Due to this, the injection amount and injection timing of the main fuelinjection are optimized and the combustion of the engine is maintainedin the optimal state.

(2) Crank Angle when (dP/dθ)max Occurs

The crank angle when (dP/dθ)max occurs is used as a combustion parameterand the fuel injection timing is corrected so that the crank angle when(dP/dθ)max occurs becomes the target value.

When the fuel injected into a combustion chamber burns, the pressureinside the combustion chamber rises. Therefore, the value of the rate ofchange (dP/dθ) of the combustion chamber pressure increases each timefuel is injected due to the main fuel injection, multi-fuel injection,etc. The same number of peaks (dP/dθ) as the injections occur.Therefore, by using the crank angle where the maximum value (maximalvalue) (dP/dθ)max of the rate of change (dP/dθ) of combustion chamberpressure at the peaks occurs as a combustion parameter, it is possibleto correct the injection timings of the fuel injections corresponding tothe fuel injection modes.

For example, when main fuel injection and multi-fuel injection areperformed, fuel is injected a plurality of times for one stroke cycle.In this case, a maximal value should occur for the combustion pressureitself corresponding to the combustion of the fuel injected by the fuelinjections, but the rise in pressure in a combustion chamber is due tothe different injections interfering with each other, so it is difficultto isolate and detect the maximal value (Pmax) for each injection.

As opposed to this, clear peak values of the rate of change of pressure(dP/dθ) appear with respect to the fuel injections. Therefore, by usingthe crank angle where (dP/dθ)max occurs as the combustion parameter, itis possible to accurately correct the injection timings of theinjections even when performing multi-fuel injection.

FIG. 3 and FIG. 4 are views for explaining the selective use of the(dP/dθ)max corresponding to the injection mode. For example, in anengine where main fuel injection, pilot injection, and after injectionare performed in one cycle, three peak values (dP/dθ)max of the rate ofchange of pressure are generated corresponding to the differentinjections. The (dP/dθ)max, (dP/dθ)No. 2max, and (dP/dθ)No. 3max in FIG.3 signify the initial, second, and third peak values among the peakvalues of the rate of change of pressure occurring after the start ofthe compression stroke (expansion stroke from end of compression stroke)(see FIG. 4).

The first column in FIG. 3 shows the fuel injection mode. In the presentembodiment, the main fuel injection is combined with multi-fuelinjection of one pilot injection and/or after injection each, so fourfuel injection modes can be conceived of: only main fuel injection,pilot injection+main fuel injection, main fuel injection+afterinjection, and pilot injection+main fuel injection+after injection.

For example, as shown in FIG. 3, when only main fuel injection isperformed, the crank angle where (dP/dθ)max occurs is used to correctthe injection timing of the main fuel injection.

Further, when the injection mode is the pilot injection+main fuelinjection, the crank angle where (dP/dθ)max occurs is used to correctthe injection timing of the pilot fuel injection, while the crank anglewhere (dP/dθ)No. 2max occurs is used to correct the injection timing ofthe main fuel injection.

Further, when the injection mode is the pilot injection+main fuelinjection+after injection, the crank angles where (dP/dθ)max, (dP/dθ)No.2max, and (dP/dθ)No. 3max occur are used to correct the injectiontimings of the pilot fuel injection, main fuel injection, and after fuelinjection.

(3) Crank Angle where Maximum Value (d2P/dθ2)max of Second Derivative ofCombustion Chamber Pressure Occurs

There is also a maximal value (d2P/dθ2)max of the second derivative ofthe combustion chamber pressure for each injection like with the maximalvalue of the rate of change of pressure explained above. Therefore, byfeedback control of the fuel injection amount and injection timing foreach injection so that the crank angle where (d2P/dθ2)max occurs becomesthe optimal value set in advance in accordance with the engine operatingconditions, it is possible to maintain the engine combustion state atthe optimal state.

FIG. 5 is a view explaining the selective use of the (d2P/dθ2)max inaccordance with the injection mode in the same way as FIG. 3. As shownin FIG. 5, (d2P/dθ2)max can be selectively used in exactly the same wayas (dP/dθ)max. Further, for example, when the change of the crank angleof (dP/dθ)max when changing the fuel injection timing is gentle, it iseffective if using (d2P/dθ2)max for correcting the injection timinginstead of (dP/dθ)max.

(4) Maximum Value Pvmax of Product of Combustion Chamber Pressure P andCombustion Chamber Actual Volume V and Crank Angle θpvmax where thatMaximum Value Occurs

PVmax has a strong correlation with the amount of cylinder heat releaseexplained later. Good precision control becomes possible in particularwhen the fuel injection mode is only main fuel injection. Further, thecombustion chamber actual volume V can be calculated in advance andstored as a function of the crank angle, so it is possible to reduce theload of calculation of the ECU 20 compared with the case of using theamount of cylinder heat release.

In the present embodiment, when performing only main fuel injection, themain fuel injection amount is feedback controlled so that the value ofPVmax becomes the optimal value set in accordance with the engineoperating conditions, while the main fuel injection timing is feedbackcontrolled so that the crank angle θpvmax where PVmax occurs becomes theoptimal value set in accordance with the engine operating conditions.

(5) Difference ΔPVmax (=PVmax−PVmaxbase) Between the Above PVmax andProduct PVmaxbase Between Combustion Chamber Pressure Due Only toCompression in Case of Assuming that no Combustion has Occurred andCombustion Chamber Actual Volume at Crank Angle θpvmax where PVmaxOccurs

FIG. 7 is a view explaining ΔPVmax=PVmax−PVmaxbase in the same way asFIG. 2. The solid line in FIG. 7 shows the change of the PV value in thecase of only main fuel injection. The product PV of the combustionchamber pressure P and the combustion chamber actual volume V, as shownin FIG. 7, greatly increases due to the rise in pressure when combustionoccurs and becomes the maximum value PVmax after top dead center TDC.

On the other hand, the broken line in FIG. 7 shows the change of the PVvalue in the case of assuming no combustion occurred, that is, the PVvalue (PVbase) calculated using the combustion chamber pressureoccurring due to only compression (motoring pressure). PVmaxbase iscalculated as the value of PVbase at the crank angle where PVmax occurs.

ΔPVmax can be used for correcting the total fuel injection amount(overall fuel injection amount) in all fuel injection modes—not onlywhen the fuel injection mode is only the main fuel injection, but alsowhen there is pilot injection or after injection. In the presentembodiment, the timing of each fuel injection is corrected so that thecrank angle where (dP/dθ)max or (d2P/dθ2)max occurs matches the targetvalue.

(6) Crank Angle where Maximum Value (dQ/dθ)max of Cylinder Heat ReleaseRate Occurs

The cylinder heat release rate is expressed by the following formulausing the cylinder pressures P detected by the cylinder pressure sensors29 a to 29 d and the actual cylinder volume V:dQ/dθ=(κ·P·(dV/dθ)+V(dP/dθ))/(κ−1)

-   -   where P and V are functions of θ, and κ expresses the specific        heat ratio of the air-fuel mixture.

The value of the cylinder heat release rate (dQ/dθ) increases each timefuel injected by the main fuel injection, multi-fuel injection, etc. isburned. In the same way as the case of the above-mentioned (dP/dθ), thesame number of peaks as the injections occur. Therefore, by using thecrank angle where each peak (dQ/dθ) occurs as a combustion parameter, itis possible to correct the fuel injection timings in accordance with thefuel injection modes.

The ECU 20 uses the crank angle detected by the crank angle sensor 25,the actual volume at the crank angle, and the cylinder pressures of thecylinders detected by the cylinder pressure sensors 29 a to 29 d tocalculate the difference for each constant crank angle, finds theapproximate value of (dQ/dθ) for each crank angle, and finds the crankangle where the (dQ/dθ)max of the maximal value of the values of (dQ/dθ)calculated occurs.

As explained above, there are the same number of (dQ/dθ)max as thenumber of fuel injections per cycle in the same way as the case of(dP/dθ)max.

FIG. 6 shows the selective use of the (dQ/dθ)max in accordance with thefuel injection mode. In FIG. 6, (dQ/dθ)max, (dQ/dθ)No. 2max, and(dQ/dθ)No. 3max signify the first, second, and third maximal values of(dQ/dθ). The selective use of the (dQ/dθ)max of FIG. 6 is exactly thesame as the case of the (dP/dθ)max of FIG. 3, so a detailed explanationwill be omitted here.

(7) Overall Amount of Cylinder Heat Release ΣdQ

The overall amount of cylinder heat release ΣdQ is found by cumulativelyadding the values of the above (dQ/dθ) for one stroke cycle. ΣdQcorresponds to the overall amount of fuel fed to a combustion chamber,so for example can be used for correction of the overall fuel injectionamount not only in main fuel injection alone, but also in a fuelinjection mode including pilot injection or after injection. Note thatthe injection timing is corrected by any of the methods explainedseparately above.

(8) Difference Pmax−Pmin Between Maximum Value Pmax of Cylinder Pressureafter Start of Combustion and Cylinder Minimum Pressure Pmin in Intervalafter Compression Top Dead Center until Combustion is Started inCombustion Chamber

FIG. 8 is a view showing Pmax and Pmin similar to FIG. 2. In this case,the main fuel injection amount is feedback controlled so that the valueof Pmax−Pmin becomes the target value, while the main fuel injectiontiming is controlled so that the crank angle θpmax where Pmax occursbecomes the target value. Pmax−Pmin is suitable for the case where thefuel injection mode is only the main fuel injection.

(9) Difference Pmax−Pmaxbase Between Maximum Value Pmax of CylinderPressure after Start of Combustion and Combustion Chamber Pressure(Motoring Pressure) Pmaxbase Due Only to Compression in Case of AssumingNo Combustion Occurred at Crank Angle where Pmax Occurs

FIG. 9 is a view showing Pmax and Pmaxbase similar to FIG. 2. As shownin FIG. 9, Pmaxbase is the change in the combustion chamber pressure inthe case of assuming no combustion occurs at the crank angle θpmax wherePmax occurs, that is, the combustion chamber pressure occurring due toonly compression (motoring pressure). Pmaxbase can be found bycalculation, but as shown in FIG. 9, the motoring pressure issymmetrical right-left with respect to compression top dead center, soin the present embodiment, the pressure detected by the cylinderpressure sensor at the compression stroke crank angle θpmax′ symmetricalwith θpmax about the compression top dead center TDC is used asPmaxbase.

Pmax−Pmaxbase is used to correct the fuel injection amount of the mainfuel injection when the fuel injection mode is only the main fuelinjection in the same way as the above Pmax−Pmin, but is particularlysuitable for correction of the fuel injection amount in the case wherethe minimum pressure Pmin does not appear after top dead center in thechange of the combustion chamber pressure as shown in FIG. 9. Further,in the present embodiment as well, the injection timing of the main fuelinjection is corrected so that the crank angle θpmax where Pmax occursmatches the target value.

(10) PVmain−Pmainbase

FIG. 10 is a view for explaining the combustion parameterPVmain−PVmainbase. In FIG. 10, the abscissa shows the crank angle, whilethe ordinate shows the product PV of the combustion chamber pressure Pand the combustion chamber actual volume V at the different crankangles.

FIG. 10 shows the case where pilot injection is performed in addition tomain fuel injection.

As shown in FIG. 10, the PV value rapidly rises when the fuel injectedby the pilot injection is ignited (FIG. 10, point P) and when the fuelinjected by the main fuel injection is ignited (same, point M).

As shown in FIG. 10, PVmain is the FV value when the fuel of the mainfuel injection is ignited (point M). Further, PVmainbase is the productof the combustion chamber pressure (motoring pressure) P obtained byonly compression and the actual combustion chamber volume V at the crankangle where the fuel of the main fuel injection is ignited (point M).

The value of PVmain can be easily detected as the point where the secondderivative of the PV value becomes positive. The motoring pressurePmainbase is found from the crank angle at that time. This Pmainbase andthe combustion chamber actual volume V are used to calculate PVmainbase.

PVmain−PVmainbase is particularly suitable for correction of the pilotinjection amount in the case of performing pilot injection.

(11) ΔPVmax−ΔPVafter

FIG. 14 is a view explaining ΔPVmax−ΔPVafter similar to FIG. 10.

FIG. 14 shows the case where pilot injection and after injection areperformed in addition to main fuel injection. Therefore, in FIG. 14,there are three points where the PV value rapidly increases (FIG. 14,points P, M, and A). Further, the crank angle where the maximum value ofPV, that is, PVmax, occurs becomes the point of time where afterinjection is performed after the start of combustion of the fuel of themain fuel injection.

As explained above, ΔPVmax is found as the difference between themaximum value of PV, that is, PVmax, and the value of PV at the time ofmotoring at a crank angle where PVmax occurs, that is, PVmaxbase.

Further, ΔPVafter is found as the difference between the product PVafterof the combustion chamber pressure detected by the cylinder pressuresensor when the fuel due to the after injection starts to burn, that is,at the point of the start of a rapid increase in the PV value (point A)occurring third after the start of combustion in FIG. 14, and thecombustion chamber actual volume at that time and the value of PV at thetime of motoring at the crank angle of point A, that is, PVafterbase.

That is,ΔPVmax=PVmax−PVmaxbaseΔPVafter=PVafter−PVafterbaseThe combustion parameter ΔPVmax−ΔPVafter is particularly suitable forcorrection of the after injection amount in the case of performing afterinjection.

(12) Difference Ptdc−Pmin of Combustion Chamber Pressure Pmtdc at TopDead Center of Compression Stroke and Combustion Chamber MinimumPressure Pmin in Interval after Compression Top Dead Center to WhenCombustion is Started in Combustion Chamber

FIG. 11 is a view showing the change in the combustion chamber pressurein the case of maintaining the fuel injection amount constant whileadjusting the amount of EGR gas to change the combustion air-fuel ratioin the same way as FIG. 2. When maintaining the fuel injection amountconstant, the combustion pressure maximum value Pmax will not changemuch at all regardless of a change in the air-fuel ratio, but the lowerthe air-fuel ratio (the higher the EGR rate), the further the ignitiontiming of the air-fuel mixture from compression top dead center. Thedifference between the combustion chamber pressure at compression topdead center and the combustion chamber minimum pressure Pmin in theperiod from top dead center to the start of combustion changes inaccordance with the air-fuel ratio.

For this reason, by using as the combustion parameter the differencePmtdc−Pmin between the combustion chamber pressure Pmtdc due tocompression at the compression stroke top dead center and the minimumpressure Pmin after top dead center and controlling the amount of EGR(for example, the throttle valve opening degree) so that this valuebecomes the target value, the combustion air-fuel ratio can be optimallycontrolled. Further, in this case, the main fuel injection amount andmain fuel injection timing are controlled using Pmax and the crank angleat which Pmax occurs.

Note that Pmtdc has to be the pressure in the state where no combustionoccurs at the compression stroke top dead center, that is, the pressuredue to only compression in the cylinder. Therefore, to prevent errorfrom occurring when combustion etc. due to the pilot injection isstarted, the value calculated by the following formula is used as Pmtdc;Pmtdc=Pbdc−·(ε)^(κ) =Pm·(ε)^(κ)

-   -   where Pbdc is the combustion chamber pressure at the suction        stroke bottom dead center and is substantially equal to the        intake pipe pressure (supercharging pressure) Pm. Further, ε is        the compression ratio of a cylinder, while κ is the specific        heat ratio of the air-fuel mixture and is preferably found by        experiments.

As explained later, control using the two combustion parameters of Pmaxand Pmtdc−Pmin is particularly effective at the transient times such aswhen switching combustion modes between normal combustion and lowtemperature combustion.

Note that it is learned that the combustion parameters shown below mayalso be used in addition to the above combustion parameters:

-   -   (a) Pmax−Pmtdc

Pmax and Pmtdc (see FIG. 11) were explained above, but Pmax−Pmtdc can beused for correction of the overall fuel injection amount (main fuelinjection amount) when only main fuel injection is performed.

(b) PVmain

PVmain was explained for the case of use of PVmain−PVmainbase as thecombustion parameter in FIG. 10, but PVmain expresses the amount of heatin a cylinder right before the fuel of the main fuel injection ignites,so by correcting the pilot injection amount so that PVmain becomes avalue predetermined in accordance with the operating conditions, it ispossible to control the pilot injection amount to a suitable value.

(c) Pmain−Pmainbase

Pmain and Pmainbase are the combustion chamber pressure when the fuel ofthe main fuel injection ignites and the motoring pressure at the crankangle where this Pmain occurs (see FIG. 15). Pmain−Pmainbase also issuitable for correction of the pilot injection amount in the same way asPVmain−PVmainbase.

(d) ΣdQmain

ΣdQmain is the value of the cylinder heat release rate (dQ/dθ) explainedabove cumulatively added from when the compression stroke starts to whenthe fuel due to the main fuel injection ignites (integral value).ΣdQmain corresponds to the overall amount of heat supplied to thecombustion chamber before the fuel of the main fuel injection starts toburn, so for example in a fuel injection mode including pilot injection,it corresponds to the injection amount of the pilot injection. For thisreason, by using ΣdQmain as the combustion parameter, it is possible tosuitably correct the pilot injection amount.

(e) ΣdQ−ΣdQafter

ΣdQ is the overall amount of cylinder heat release explained above,while ΣdQafter is the cumulative value (integrated value) of thecylinder heat release rate (dQ/dθ) from the start of the compressionstroke to when the fuel of the after injection ignites. ΣdQaftercorresponds to the total of the amount of heat supplied to thecombustion chamber up to when the fuel of the after injection starts toburn, so ΣdQ−ΣdQafter corresponds to the total of the amount of heatsupplied to the combustion chamber due only to after injection, that is,the fuel injection amount of the after injection. Therefore, by usingΣdQ−ΣdQafter as the combustion parameter, it is possible to suitablycorrect the injection amount of the after injection.

Next, the flow chart of FIG. 12 will be used to explain the fuelinjection correction operation of the present embodiment.

The operation of FIG. 12 is executed by the ECU 20. Below, the operationof the steps of FIG. 12 will be explained.

Step 1201:

Step 1201 shows the judgment as to whether the conditions for executionof the control operation stand. At step 1201, it is decided whether toexecute the control operation of step 1203 on based on the cumulativeoperating time of the engine or the cumulative running distance of thevehicle.

For example, when the cumulative operating time from when the vehiclestarted being operated has not reached a predetermined value (or whenthe running distance of the vehicle has not reached a predeterminedvalue), the initial friction of the engine parts is large, so ifcorrecting the injection amount or injection timing, error is likely tooccur. Accordingly, the control operation of FIG. 12 is executed onlywhen the cumulative operating time of the engine is a predetermined timeor more.

Step 1203:

At step 1203, the cylinder pressure sensors 29 a to 29 d are calibrated.Here, the deviations of the zero points (offsets) and gains of thecylinder sensors are corrected.

FIG. 13 is a view for explaining the calibration of a cylinder pressuresensor.

In FIG. 13, the abscissa shows the crank angle, while the ordinate showsthe cylinder pressure. Further, BDC of the abscissa indicates bottomdead center of the suction stroke, while TDC indicates top dead centerof the compression stroke. Further, CR is a suitable crank angle beforethe start of combustion during the compression stroke.

The solid line in FIG. 13 shows the change of the actual output of acylinder pressure sensor, while the broken line shows the change of thetrue cylinder pressure. Further, PR1 and PC1 show the output of thecylinder pressure sensor and true cylinder pressure at the suctionstroke bottom dead center BDC, while PR2 and PC2 show the output of thecylinder pressure sensor and true cylinder pressure at the crank angleCR.

Here, if the true cylinder pressures PC1 and PC2 are known, the offsetΔPR of the cylinder pressure sensor is found as ΔPR=PC1−PR1, thedeviation K of the gain is found as K=PC2/(PR2+ΔPR), and the truecylinder pressure PC when the output of the cylinder pressure sensor isPR can be found as:PC=K·(PR+ΔPR)

Here, PC1 is the cylinder pressure at the suction stroke bottom deadcenter and is approximately equal to the suction pressure (superchargingpressure) Pm. Therefore, the offset ΔPR becomes ΔPR=Pm−PR1.

Further, the true cylinder pressure PC2 at the crank angle CR iscalculated as PC2=Pm·(εcr)^(κ) using PC1=Pm. Here, εcr is thecompression ratio at the crank angle Cr, while κ is the specific heatratio of the air-fuel mixture. That is, the value of the gain deviationK is found asK=Pm·(εcr)^(κ)/(PR 2+ΔPR)

In the present embodiment, before executing the control operation, theECU 20 finds the above offset APR and gain K based on the output of thecylinder pressure sensor at the suction stroke bottom dead center andcrank angle CR and the engine supercharging pressure Pm and in thefollowing operation converts the output PR of the cylinder pressuresensor to the true cylinder pressure PC (PC=K·(PR+ΔPR)) for use.

Step 1205:

At step 1205, the engine speed Ne and the accelerator opening degreeAccp are read from the crank angle sensor 25 and the accelerator openingdegree sensor 21. Ne and Accep are used for setting the target value ofa combustion parameter explained later. Note that in the presentembodiment, the fuel injections and fuel injection timings of the mainfuel injection and multi-fuel injection are calculated based on Ne andAccp by a not shown fuel injection control operation performedseparately by the ECU 20.

Step 1207:

At step 1207, the combustion parameter giving the least error isselected based on the current fuel injection mode of the engine (onlymain fuel injection or main fuel injection+multi-fuel injection).

In the present embodiment, the value of the combustion parameterselected in accordance with the fuel injection mode from for example theabove-mentioned 11, combustion parameters is calculated. When forexample the current fuel injection mode is only the main fuel injection,Pmax, PVmax, etc. are selected as the combustion parameters, while whenfor example the current fuel injection mode is the pilot injection+mainfuel injection+after injection multi-fuel injection, PVmain−PVmainbase,ΔPVmax, ΔPVmax−ΔPVafter, etc. are selected as the combustion parametersfor correcting the fuel injection amounts, while (dP/dθ)max, (dQ/dθ)max,etc. are selected for correcting the fuel injection timings of the fuelinjections.

Step 1209 and Step 1211:

At step 1209 and step 1211, first the overall fuel injection amount andthe injection timing of the main fuel injection are corrected. That is,at step 1209, first the magnitude of the combustion parameter (forexample, ΔPVmax) selected at step 1207 is calculated based on thecylinder pressure sensor output. The overall fuel injection amount iscorrected to increase or decrease it until this ΔPVmax matches thetarget value of ΔPVmax determined from the engine speed Ne and theaccelerator opening degree Accp.

Note that the target values of the combustion parameters are found fromexperiments etc. in advance and stored as numerical value maps using Neand Accep in the ROM of the ECU 20.

Further, at step 1211, similarly the crank angle where the stateselected as the combustion parameter (for example, (dP/dθ)max) occurs isdetected based on the output of the cylinder pressure sensors, and theinjection timing of the main fuel injection is corrected until the crankangle matches with the target value determined from the speed Ne and theaccelerator opening degree Accp.

Step 1213:

Step 1213 shows the correction of the injection amount and injectiontiming of the multi-fuel injection in the case of execution ofmulti-fuel injection. At this step, for example, the injection amountand injection timing of the pilot injection and/or after injection arecorrected until PVmain−PVmainbase, ΔPVmax−ΔPVafter, (dP/dθ)max,(dP/dθ)No. 3max, and other combustion parameters match their targetvalues.

The specific correction is similar to that for the main fuel injection,so a detailed explanation will be omitted here, but in the presentembodiment, first, the overall fuel injection amount, the injectionamount and injection timing of the main fuel injection, etc. arecorrected, then the injection timing and injection amount of themulti-fuel injection (pilot injection and after injection) arecorrected. This is because even when performing multi-fuel injection,the overall fuel injection amount has the greatest effect on the outputtorque, so first the overall fuel injection amount is optimallycorrected, then the injection amount and injection timing of the mainfuel injection are optimally corrected so as to bring the combustionstate of the engine close to the ideal state, then the injection amountand injection timing of the multi-fuel injection are corrected for fineadjustment of the combustion state.

As explained above, by performing the correction operation of FIG. 12,the injection amounts and injection timings of the different fuelinjections are corrected to suitable values and the combustion state ofthe engine is optimized.

Next, another embodiment of the present invention will be explained.

In the present embodiment, the fuel injection is controlled at thetransient time of switching of combustion modes. In the presentembodiment, the engine 1 operates while switching between the twocombustion modes of a normal diesel combustion mode, that is, acombustion mode injecting fuel and burning with a high air-fuel ratio atthe end of the compression stroke, and a low temperature combustionmode, that is, a combustion mode greatly advancing the fuel injectiontiming to form a premixed air-fuel mixture in the cylinder and greatlyincreasing the amount of EGR gas to burn with a low air-fuel ratio.Further, when switching the combustion modes, a combustion parameter isused for feedback control of the fuel injection and a throttle valveprovided in the engine intake passage is feedback controlled to adjustthe amount of intake air and optimize the air-fuel ratio.

As will be explained in FIG. 11, there is a good correlation between(Pmtdc−Pmin) among the combustion parameters and the air-fuel ratio.When switching between the normal diesel combustion mode and lowtemperature combustion mode, the air-fuel ratio and fuel injectiontiming greatly change. Here, the air-fuel ratio is adjusted by adjustingthe opening degree of the throttle valve to change the amount of EGRgas, but the change of the amount of EGR gas takes a relatively longtime, while the fuel injection timing can be changed in a short time.Therefore, in the present embodiment, at the time of switching, first,Pmax is used as a combustion parameter to correct the fuel injectionamount, then the throttle valve opening degree is used as a combustionparameter (Pmtdc−Pmin) for correction, then the crank angle where Pmaxoccurs after correction of the throttle valve opening degree is used asa combustion parameter to correct the fuel injection timing.

The amount of EGR gas, which is inherently slow in speed of change, isfirst corrected and then the injection timing is corrected in this waybecause at the start of the switch, generally the change of thecombustion parameter becomes small and the sensitivity becomes lowrelative to the change of the fuel injection timing and so as to preventthe problem of dispersion of control in the case of simultaneouslycontrolling the air-fuel ratio and injection timing.

FIG. 16 is a flow chart showing an outline of the combustion modeswitching control operation. This operation is executed by the ECU 20.

In the operation of FIG. 16, at step 1601, first, the engine speed Neand accelerator opening degree Accp are read from the crank angle sensor25 and accelerator opening degree sensor 21, then at step 1603, Pmax iscalculated based on the cylinder pressure sensor output. Further, atstep 1605, the fuel injection amount is feedback controlled until thevalue of this Pmax matches the Pmax target value determined from theengine speed Ne and the accelerator opening degree Accp.

Further, after the feedback control of the fuel injection amount of step1605 is completed, at step 1607, the parameter (Pmtdc−Pmin) iscalculated based on the cylinder pressure sensor output and at step1609, the throttle valve opening degree is feedback controlled until thevalue of (Pmtdc−Pmin) matches the target value determined from theengine speed Ne and the accelerator opening degree Accp.

Further, after the throttle valve opening degree finishes being adjustedat step 1609, at step 1611, it is again judged if the valve of Pmax isconverging on the target value. When Pmax is away from the target valueby a predetermined amount or more, the operation from step 1601 isexecuted again.

When Pmax is converging on the target value at step 1611, the routinenext proceeds to step 1613 where the crank angle where Pmax occurs iscalculated from the cylinder pressure sensor output and the fuelinjection timing is feedback controlled until the crank angle matcheswith the target value determined from the engine speed Ne and theaccelerator opening degree Accp.

By the switching control of FIG. 16, even during transient operationsuch as combustion mode switching, the engine combustion state isoptimally controlled.

Next, control of the amount of EGR gas based on a combustion parameterwill be explained in detail. In the present embodiment, as combustionparameters, PVmax, θpvmax, ΔPVmax, and Δt are used for feedback controlof the amount of EGR gas, the fuel injection amount, and the fuelinjection timing.

FIG. 17 shows the combustion parameters used in the present embodiment,that is, PVmax, θpvmax, ΔPVmax, and Δt.

In FIG. 17, the abscissa shows the crank angle (CA) from the compressionstroke to expansion stroke of a cylinder, while the ordinate shows theabove-mentioned PV value. On the abscissa, TDC shows the compression topdead center.

The PV value is the product of the pressure and volume, so becomes avalue corresponding to the cylinder temperature due to the relationshipPV≠MRT of the equation of state of a gas (where M: number of moles ofgas, R: general gas constant (J/mol·K), and T: temperature (° K)).Further, the timing when PV becomes the maximum value PVmax (FIG. 17,θpvmax) is confirmed by experiments to correspond to the point of timewhen the fuel injected into a cylinder finishes being burned (strictlyspeaking, the point of time when 90% of the fuel is burned). Therefore,θpvmax can be used as an indicator expressing the time of end ofcombustion in a cylinder.

In FIG. 17, θinj shows the fuel injection start timing from a fuelinjector (10 a to 10 d, hereinafter referred to generally by thereference numeral 10). Further, in FIG. 17, Δt shows the combustioncompletion time defined as the time (crank angle) from the fuelinjection start (θinj) to the combustion end (θpvmax). The fuel injectedfrom a fuel injector 10 starts to burn after the elapse of a certainignition delay time and finishes burning after the elapse of acombustion time determined by various conditions. Therefore, thecombustion completion time Δt (=θpvmax−θinj) corresponds to the total ofthe ignition delay time and combustion time of the fuel.

Further, in FIG. 17, the broken line shows the change (Pvbase) of the PVvalue when no combustion occurs in the cylinder. PVbase expresses thecompression and expansion of gas in a cylinder due to the up and downmotion of the piston, so becomes a curve symmetric about top deadcenter.

As explained above, ΔPVmax is defined as the difference between themaximum value PVmax of the PV value and the value PVmaxbase of PVbase atθpvmax.

The value PVmaxbase of PVbase at θpvmax can be easily calculated fromthe cylinder pressure at the end of the suction stroke and the cylindervolume at θpvmax. However, as explained above, the PVbase curve becomessymmetric about compression top dead center. Therefore, in the presentembodiment, after detection of θpvmax, the value of PVbase at the pointof the compression stroke becoming symmetric about top dead center(shown by θpvmax′ in FIG. 17) is used to calculate ΔPVmax, but inpractice the PV value and PVbase value become identical in thecompression stroke where combustion occurs. For this reason, in thepresent embodiment, by actually using the PV value at θpvmax′ as the PVbase value at θpvmax, the value of ΔPVmax is easily calculated.

Next, the meanings of the combustion parameters Δt, PVmax, θpvmax, andΔPVmax will be explained.

As explained above, the period from the start of fuel injection toθpvmax, that is, the combustion completion time Δt, corresponds to thetotal of the ignition delay time and combustion time of the injectedfuel. On the other hand, the ignition delay time and combustion time areboth greatly affected by the EGR rate (ratio of amount of EGR gas in gassucked into cylinder). As the EGR rate becomes larger, the Δt alsoincreases. Therefore, the combustion completion time Δt has a closecorrelation with the cylinder EGR rate and can be used as an indicatorexpressing the EGR rate.

Further, the timing θpvmax where PVmax occurs is correlated with the endtiming of the combustion and is greatly related to the combustion statein the cylinder. Further, if the other conditions are the same, the endtiming of the combustion changes according to the fuel injection timing.

Further, the value of ΔPVmax is the difference (temperature difference)between the PV values at the time of combustion and the time whencombustion does not occur, so is correlated with the amount of fuelburned in the combustion chamber, that is, the fuel injection amount.

The present embodiment takes note of the above and uses the Δt, θpvmax,and ΔPVmax for feedback control of the amount of EGR gas, fuel injectiontiming, and fuel injection amount to their optimum values.

That is, in the present embodiment, the engine is operated in advancewhile changing the engine operating conditions (accelerator openingdegree and speed in combination) so as to search for the fuel injectionamount, fuel injection timing, and EGR rate (EGR valve opening degree)giving the optimum combustion state in terms of the fuel efficiency,exhaust gas properties, etc. and these values are used as referencevalues for the fuel injection amount, fuel injection timing, and EGRvalve opening degree at the different operating conditions and stored inthe form of a two-dimensional numerical value map using the acceleratoropening degree and speed (hereinafter referred to for convenience as the“reference injection condition map”) in the ROM of the ECU 20.

Further, in the present embodiment, the values of the combustionparameters Δt, θpvmax, and ΔPVmax at the time of giving the optimumcombustion state at the above different operating conditions arecalculated and stored in the form of a two-dimensional numerical valuemap using the accelerator opening degree and speed (hereinafter referredto for convenience as a “target characteristic map”) in the ROM of theECU 20.

In actual operation, the ECU 20 first uses the above reference injectioncondition map to find the fuel injection amount, fuel injection timing,and EGR valve opening degree from the engine speed and acceleratoropening degree and controls the fuel injection amount, fuel injectiontiming, and EGR valve opening degree to the reference injectioncondition map values.

Further, in this state, it calculates the combustion parameters Δt,θpvmax, and ΔPVmax of each cylinder based on the output from thecylinder pressure sensors 29 a to 29 d. Further, it uses the currentaccelerator opening degree and speed to find the target values Δt,θpvmax, and ΔPVmax of the combustion parameters at the optimumcombustion state from the above-mentioned target characteristic map andadjusts the fuel injection amount, fuel injection timing, EGR valveopening degree, etc. determined from the reference injection conditionmap so that the actual combustion parameters match their target values.

Specifically, the ECU 20 adjusts the opening degree of the EGR valve 35for feedback control so that the actual combustion parameter Δt becomesthe target value and feedback controls the fuel injection timing andfuel injection amount so that the θpvmax and ΔPVmax match their targetvalues.

Due to this, the EGR and fuel injection are controlled so that theactual combustion state becomes the optimum state.

FIG. 18 and FIG. 19 are flow charts for specifically explaining thecontrol operation based on the above combustion pressure characteristics(combustion parameter control operation). The operations of FIG. 18 andFIG. 19 are performed as routines executed by the ECU 20 every constantinterval.

FIG. 18 shows the basic control operation of fuel injection and EGR. Inthe operation of FIG. 18, the ECU 20 sets the fuel injection amount,fuel injection timing, and EGR valve 35 opening degree as the sum of thereference values determined from the engine speed Ne and acceleratoropening degree Accp and the correction amounts determined based on thecombustion parameters from the operation of FIG. 19.

In FIG. 18, at step 301, the accelerator opening degree Accp and theengine speed Ne are read. At step 303, the values of Accp and Ne read atstep 301 are used to read the reference fuel injection amount FI₀,reference fuel injection timing θI₀, and reference EGR valve openingdegree EGV₀ from the above-mentioned reference injection condition mapstored in the ROM of the ECU 20 in advance in the form a two-dimensionalnumerical value map using Accp and Ne.

The reference fuel injection amount, reference fuel injection timing,and reference EGR valve opening degree are the fuel injection amount,fuel injection timing, and EGR valve opening degree giving the optimumcombustion state as found by actually operating the engine in advance.

The above reference values are the fuel injection amount and timing andEGR valve opening degree able to give the optimum combustion state inthe environment at the time of experiments, but in actual operation,there are differences in fuel, differences in the engine operatingenvironment (air temperature, atmospheric pressure, etc.), variations inequipment, changes in characteristics, etc., so even if operating usingthese reference values, the optimum combustion state is not necessarilyobtained.

Therefore, in the present embodiment, the values of the reference valuesFI₀, θI₀, and EGV₀ found as explained above to which the correctionamounts α, β, and γ are added for correction are set as the actual fuelinjection amount, fuel injection timing, and EGR valve opening degree.That is, at step 305, the actual fuel injection amount FI, the fuelinjection timing θI, and the EGR valve opening degree EGV are set asFI=FI₀+α, OI=θI₀+β, and EGV=EGV₀+γ and at step 307, the values set atstep 305 are used for control of the fuel injection and EGR valveopening degree.

Here, α, β, and γ are feedback correction amounts set based on thecombustion parameters by the operation of FIG. 19.

Explaining the operation of FIG. 19, first, at step 401, the acceleratoropening degree Accp and the engine speed Ne are read. Further, at step403, the target values θpvmax₀, ΔPVmax₀, and Δt₀ of the θpvmax, ΔPVmax,and Δt are read from the two-dimensional map using Accp and Ne stored inthe ROM of the ECU 20 in advance. The target values θpvmax₀, ΔPVmax₀,and Δt₀ are the values of θpvmax, ΔPVmax, and Δt when the optimumcombustion is obtained at those accelerator opening degree and speed.

Further, at step 405, the combustion parameters θpvmax, ΔPVmax, and Δtof the cylinders are calculated based on the outputs of the cylinderpressure sensors 29 a to 29 d.

Further, from steps 407 to 411, the correction amounts α, β, and γ arefeedback corrected so that the values of the actual combustionparameters calculated at step 405 match with the target values foundfrom the map at step 403.

That is, at step 407, first, the correction amount a of the fuelinjection amount is feedback controlled so that the actual value ofΔPVmax matches with the target value ΔPVmax₀, at step 409, thecorrection amount β of the fuel injection timing is feedback controlledso that the actual value of θpvmax matches with the target valueθpvmax₀, and at step 411, the correction amount γ of the EGR valveopening degree is feedback controlled so that the actual value of Δtmatches the target value Δt₀. The feedback control from step 407 to 411is for example made PID control based on the deviation of the actualvalues from the target values.

For example, specifically explaining the PID control in the presentembodiment using as an example the correction amount β of the fuelinjection timing, if the deviation between the actual value of θpvmaxand the target value θpvmax₀ is δ, the correction amount β is calculatedusing the following equation:β=K ₁ ×δ+K ₂ ×Σδ+K ₃×(δ−δ_(i-1))

-   -   where the first term K₁×δ at the right side is a proportional        term, the second term K₂×Σδ is an integration term, and Σδ        expresses the cumulative value of the deviation δ (integrated        value). Further, the third term K₃×(δ−δ_(i-1)) is a derivative        term, while (δ−δ_(i-1)) expresses the amount of change of the        deviation δ from the previous time (derivative) (δi−1 is the        value of δ of the previous time). Further, K₁, K₂, and K₃ are        constants.

As explained above, by repeating the operations of FIG. 18 and FIG. 19,the actual fuel injection amount, fuel injection timing, and EGR valveopening degree (EGR rate) are controlled so that the combustionparameters match the target values.

By feedback control of the fuel injection amount, fuel injection timing,and EGR rate so that the combustion parameters match the target valuesin actual operation in this way, it is possible to easily obtain theoptimum combustion state without individually considering for examplethe differences in the operating environment of the engine, changes incharacteristics or variations in the equipment, differences in fuel,etc.

Note that in the operations of FIG. 18 and FIG. 19, by controlling thefuel injection amount and timing etc. to the reference values andfeedback controlling the correction amounts for these references usingcombustion parameters, the fuel injection amount etc. are converged withthe values giving the optimum combustion state in a short time. However,it is also possible to feedback control the fuel injection amount andtiming and the EGR rate themselves using the combustion parameterswithout presetting reference values of the fuel injection amount etc.

Note that when controlling the fuel injection timing θI based on thedeviation δ between the θpvmax and θpvmax₀ as in FIG. 18 and FIG. 19, inparticular when the target value of the fuel injection timing itself isgreatly advanced such as during operation in the above-mentioned lowtemperature combustion mode, the control may become dispersed.

For example, when the actual value of θpvmax is delayed from the targetvalue θpvmax₀, the fuel injection timing θI is advanced to make theθpvmax faster. However, when the fuel injection timing is already setgreatly advanced like in low temperature combustion, if overly advancingthe fuel injection timing, the combustion becomes unstable and misfireseasily occur, so if advancing the fuel injection timing, sometimes theθpvmax conversely will become slower.

In such a case, if using the θpvmax to control the fuel injectiontiming, the fuel injection timing will end up being further advanced andnot only will the control become dispersed, but also the excessiveadvance of the fuel injection will cause fuel to be injected at aposition where the piston has not sufficiently risen in the cylinder,the injected fuel will overflow from the inside to the outside of thebowl formed on the top of the piston or the injected fuel will directlystrike the cylinder wall (bore flushing) and liquid fuel will deposit onthe cylinder wall, so the problems will arise of dilution of thelubrication oil or deterioration of the fuel efficiency and exhaustproperties.

In particular, when simultaneously using Δt to control the amount of EGRgas like at step 411 of FIG. 19, if the fuel injection timing isexcessively advanced, the value of Δt will also become excessive and theamount of EGR gas will be greatly reduced, so the change of the fuelinjection timing and the adjustment of the amount of EGR gas will affecteach other and control will not longer be stable in some cases.

Therefore, in the present embodiment, an advance guard value θImax isprovided for the fuel injection timing θI calculated at step 305 so thatthe fuel injection timing will not advance more than θImax.

Specifically, if the fuel injection timing θI is calculated as θI=θI₀+βat step 305 of FIG. 18, the ECU 20 compares the calculated θI and theadvance guard value θImax. When θI is set advanced by θImax or more(θI≧θImax), it uses the θImax instead of the calculated θI to executethe fuel injection control at step 307. That is, the value of θIcalculated at step 305 is used at step 307 only when at the delayed sidefrom the advance guard value θImax (θI≦θImax).

Due to this, excessive advance in feedback control of the fuel injectiontiming using the combustion parameter θpvmax is prevented, so dilutionof the lubrication oil and deterioration of the fuel efficiency andexhaust properties due to bore flushing are prevented, dispersion offuel injection timing control due to excessive advance and interferencewith feedback control of the amount of EGR gas using Δt are prevented,and the fuel injection timing or amount of EGR gas converge to theirtarget values in a short time.

Note that the advance guard value θImax or the fuel injection timing isthe timing by which the fuel injected from a fuel injector will notoverflow to the outside from the inside of the bowl of the piston ordeposit on the wall and is a value determined by the engine speed andthe fuel injection pressure and other injection conditions. This valuediffers depending on various conditions such as the shape of the piston,the arrangement of the fuel injector, the engine speed, and theinjection pressure, so it is desirable to prepare it as a numericalvalue map for each speed (fuel injection pressure) based on experimentsusing actual engines.

Next, another embodiment of the present invention will be explained.

In the present embodiment, the engine 1 operates switching between thetwo combustion modes of the normal diesel combustion mode, that is, thecombustion mode injecting fuel at the end of the compression stroke anddiffusing and burning it with a high air-fuel ratio, and the lowtemperature combustion mode that is, the combustion mode greatlyadvancing the fuel injection timing to form a premixed air-fuel mixturein the cylinder and greatly increasing the amount of EGR gas to burn thefuel with a low air-fuel ratio. In low temperature combustion, whilecombustion with a relatively low air-fuel ratio, a large amount of EGRgas is supplied to a combustion chamber so as to greatly suppress thegeneration of NO_(x) and other harmful substances. Further, while adiesel engine, premix combustion may be performed so as to greatlyreduce the amount of generation of soot.

However, in operation in the low temperature combustion mode, thecombustion state changes extremely sensitively with respect to changesin the EGR rate. In some cases, the combustion state will greatlydeteriorate with even just a small change of the EGR rate.

Therefore, in the present embodiment, when the engine is operated in thelow temperature combustion mode, the EGR rate (EGR valve opening degree)is feedback controlled based on a combustion parameter.

FIG. 20 is a flow chart for explaining the EGR rate control operationusing a combustion parameter of the present embodiment.

In the operation of FIG. 20, first, at step 501, it is judged if theengine is currently being operated in the low temperature combustionmode. When not being operated in the low temperature combustion mode,the operation is ended immediately without executing step 503 on. Inthis case, for example, the EGR rate is controlled by open loop controlbased on the accelerator opening degree and the engine speed in the sameway as the past.

When the engine is currently being operated in the low temperaturecombustion mode at step 501, the routine next proceeds to step 503,where the current accelerator opening degree Accp and the engine speedNe are read from the corresponding sensors. At step 505, the targetvalue Δt₀ of Δt at the current Accp and Ne is read from the target valuemap of the combustion completion time Δt stored in advance in the ROM ofthe ECU 20 in the form of a two-dimensional numerical value map of Accpand Ne.

Here, Δt₀ is the combustion completion time when supplying EGR gate byan EGR rate giving the optimum combustion state in the low temperaturecombustion mode.

Next, at step 507, the current actual combustion completion time Δt iscalculated based on the outputs of the cylinder pressure sensors 29 a to29 d. Further, at step 509, the EGR valve opening degree is feedbackcontrolled so that the actual combustion completion time Δt matches withthe target value Δt₀. This feedback control, like the case of FIG. 19,is for example made PID control based on the deviation between thetarget value Δt₀ and actual value Δt.

Note that in the present embodiment, the fuel injection amount and fuelinjection timing are set to optimum values for operation in the lowtemperature combustion mode in advance by a routine executed separatelyby the ECU 20.

As shown in FIG. 20, in particular at the time of operation in the lowtemperature combustion mode sensitive to change of the EGR rate, it ispossible to control the EGR rate of the engine based on the combustionparameter Δt so as to obtain a stable, optimum combustion state even atthe time of low temperature combustion. However, as explained above,control based on Δt enables the optimum EGR rate to be obtained aftertransition to the low temperature combustion mode, but when shiftingfrom the normal combustion mode to the low temperature combustion mode,if using feedback control based on Δt to adjust the amount of EGR gas,sometimes a relatively long time will be taken until convergence to theEGR rate of after transition to the low temperature combustion mode.

As explained above, in the low temperature combustion mode, the fuelinjection timing is greatly advanced compared with the normal combustionmode. However, at the time of transition to the low temperaturecombustion mode, if advancing the fuel injection timing all at once, therapid change in the combustion state will cause the engine output torqueto fluctuate and cause so-called “torque shock” to occur. Therefore,when shifting from the normal combustion mode to the low temperaturecombustion mode, a certain transition period is provided and during thistransition period (time) the fuel injection timing is relatively gentlycontinuously changed from the value in the normal combustion mode to thetarget value in the low temperature combustion mode, i.e., transitionalprocessing is performed.

Therefore, during the transitional processing, the fuel injection timingused for calculation of Δt (FIG. 17, θinj) is made to gradually change(advanced) and in accordance with this, the timing where PVmax (FIG. 17,θpvmax) occurs is also gradually changed (advanced), so at the start ofthe switch, the value of Δt will not change that much from the valuebefore the switch and will be a relatively small value.

Therefore, the difference between the target value Δt₀ of Δt afterswitching to the low temperature combustion mode and the actual Δt alsowill not become that much greater compared with before the switch. Thechange in the opening degree of the EGR value will also becomerelatively small. That is, the opening degree of the EGR valve willgradually change along with the advance of the fuel injection timing.The change in the actual amount of the EGR gas becomes slower than thechange of the EGR valve opening degree, so at the time of switching, ifthe change of the opening degree of the EGR valve is small, the changein the actual amount of EGR gas will end up becoming considerablygentle.

On the other hand, in the low temperature combustion mode, the amount ofEGR gas has to be greatly increased compared with the normal combustionmode, but as explained above, if gradually changing the opening degreeof the EGR valve in accordance with the actual Δt during the transitionperiod, due to the delay in the change of the amount of the EGR gas, theEGR rate will not reach the target value of the low temperaturecombustion mode even at the time of completion of the transition to thelow temperature combustion mode (when fuel injection timing reachestarget value), i.e., time will be taken for convergence to the targetvalue.

Therefore, in the present embodiment, when calculating the Δt during thetransition period, the actual fuel injection timing is not used. Thetarget fuel injection timing after the completion of transition to thelow temperature combustion is used. Due to this, at the time of start ofthe transition period, the value of Δt becomes much larger than the caseof use of the actual fuel injection timing. The deviation with the Δttarget value after completion of transition also becomes large. In thepresent embodiment, the EGR valve opening degree is feedback controlledbased on the deviation between the Δt and the Δt target value, so theEGR valve opening degree also greatly changes.

FIG. 21 is a view for explaining the change in Δt in the transitionperiod when switching from the normal combustion mode to the lowtemperature combustion mode in the present embodiment.

In FIG. 21, the curve θinj shows the change in the fuel injectiontiming, while the curve θpvmax shows the change in the timing where thePVmax occurs. The actual Δt becomes equal to the distance between thetwo curves (see FIG. 21).

In FIG. 21, when the transition period of switching from the normalcombustion mode is started, the fuel injection timing θinj iscontinuously advanced and becomes the target fuel injection timing inthe low temperature combustion mode when the transition period ends.

In this case, as shown in FIG. 21, the θinj will not greatly change evenat the start of transition, so the Δt using the actual fuel injectiontiming (actual Δt) will not become that large a value from the start ofthe transition period and the opening degree of the EGR valve will alsonot greatly change. For this reason, the change in the amount of the EGRgas becomes considerably gentle and the change of the θpvmax alsobecomes gentle as shown by the solid line in FIG. 21. For this reason,the value of θpvmax will not reach the target value at the lowtemperature combustion even when the switching of the fuel injectiontiming is completed and therefore a delay time will occur as shown inFIG. 21 until the θpvmax reaches the target value (that is, until theEGR rate reaches the target value).

As opposed to this, Δt calculated using the fuel injection timing targetvalue after switching to the low temperature combustion mode instead ofthe actual fuel injection timing, as shown in FIG. 21, becomes a valuelarger than the actual Δt. Therefore, in the present embodiment, the EGRvalue opening degree also greatly changes and the speed of change(increase) of the amount of EGR gas also becomes faster, so θpvmaxchanges as shown by the broken line in FIG. 21 and the occurrence of adelay time as in the case of use of the actual Δt can be prevented.

Due to this, in the present embodiment, when switching from the normalcombustion mode to the low temperature combustion mode, it becomespossible to make the EGR rate converge to the target value in a shorttime.

Next, control at the time of switching from the low temperaturecombustion mode to the normal combustion mode opposite to the above willbe explained using FIG. 22.

For example, consider the case of controlling the fuel injection amount,fuel injection timing, EGR rate, etc. using combustion parameters onlyduring low temperature combustion mode operation and performing theconventional open loop control at the time of the normal combustionmode.

In this case, at the time of low temperature combustion mode operation,the fuel injection amount, fuel injection timing, amount of EGR gas,etc. are feedback controlled based on the combustion parameters (Δt,θpvmax, ΔPVmax, etc.) and the actual fuel injection amount, fuelinjection timing, amount of EGR gas, etc. include feedback correctionamounts.

For example, explaining this taking as an example the fuel injectiontiming, as explained at step 305 of FIG. 18, the actual fuel injectiontiming during low temperature combustion becomes the target value θI₀ towhich the feedback correction amount β is added.

Normally, as shown in FIG. 22, even when switching from the lowtemperature combustion mode to the normal combustion mode, a transitionperiod similar to that explained in FIG. 21 is provided. The targetvalue of the fuel injection timing is made to change continuously in atransition period from that of the time of the low temperaturecombustion mode to the target value at the time of the normal combustionmode.

However, as explained above, the actual fuel injection timing in the lowtemperature combustion mode includes the feedback correction amount β.The fuel injection timing in the normal combustion mode is the targetfuel injection timing (open loop control) not including a feedbackcorrection amount β. For this reason, at what point of time to stop thefeedback control and make the feedback correction amount β zero becomesan issue. For example, if stopping the feedback control immediatelyalong with the start of the transition period, the fuel injection timingwould rapidly change by the feedback correction amount β simultaneouslywith the start of the transition period and torque fluctuation mightoccur due to the rapid change in the fuel injection timing. The sameapplies in the case of continuing the feedback control during thetransition period and stopping the feedback control along with thecompletion of transition.

Therefore, in the present embodiment, as shown in FIG. 22, while thefeedback control is stopped at the same time as the start of thetransition period, the feedback correction amount β at the time of startof the transition period is not immediately made 0, but the feedbackcorrection is made to be gradually reduced continuously so that itbecomes 0 at the time of the end of the transition period.

In FIG. 22, the broken line shows the target value θI₀ of the fuelinjection timing, while the solid line shows the actual fuel injectiontiming θI. As shown in the figure, in the low temperature combustionmode operation, feedback control is performed based on the combustionparameter θpvmax. A difference of exactly the feedback correction amountβ occurs between the target θI₀ and actual fuel injection timing θI.

When the transition period is started, in the present embodiment, thefeedback control is immediately stopped, but at the start of transition,the actual fuel injection timing θI is maintained as is as the valueincluding the feedback correction amount β at the time of start of thetransition period. Therefore, in the present embodiment, sudden changeof the fuel injection timing due to the stop of the feedback control atthe time of start of the transition period is prevented.

Further, as shown in FIG. 22, during the transition period, the value ofβ is continuously reduced so as to become 0 at the end of the transitionperiod (for example, the value of β is reduced proportionally to theelapsed time after the start of the transition period). Due to this,during the transition period, the actual fuel injection timing θIgradually approaches the target fuel injection timing θI₀ and matchesθI₀ at the time of the end of the transition period. Due to this, in thepresent embodiment, it becomes possible to shift from the feedbackcontrol of the fuel injection timing during the low temperaturecombustion mode to open loop control in the normal combustion modewithout any torque fluctuation.

Note that FIG. 22 was explained using as an example the fuel injectiontiming, but needless to say the fuel injection amount or amount of EGRgas may be similarly transitionally controlled.

Next, another example of application of EGR control using a combustionparameter will be explained. In the above embodiments, the combustionparameter Δt was used to accurately control the EGR rate and enable theoptimum EGR rate to be obtained for combustion even at the time of lowtemperature combustion.

For example, when purifying exhaust by providing a known NO_(x)storing/reducing catalyst which stores NO_(x) in the exhaust byabsorption, adsorption, or both when the air-fuel ratio of the exhaustflowing into the engine exhaust passage is lean and reduces and purifiesthe stored NO_(x) using the CO and other reduction ingredients orhydrocarbons etc. in the exhaust when the air-fuel ratio of theinflowing exhaust becomes rich, a need arises to accurately control theexhaust air-fuel ratio (engine air-fuel ratio) at the time of reducingand purifying the NO_(x) stored in the NO_(x) storing/reducing catalyst.However, with the above control, while it is possible to obtain theoptimal EGR rate with a good response, it is not always possible toaccurately control the combustion air-fuel ratio (exhaust air-fuelratio) of the engine.

For example, when the injection characteristics of a fuel injectorchange due to wear of the internal mechanism etc. or when there isvariation of the injection characteristics in each product, even ifcontrolling the combustion parameter to the target value, the targetair-fuel ratio is not necessarily obtained.

On the other hand, to control the exhaust air-fuel ratio to the targetair-fuel ratio, it is also possible to arrange an air-fuel ratio sensorin the exhaust passage to directly measure the exhaust air-fuel ratioand thereby feedback control the EGR control valve.

However, EGR control using an air-fuel ratio sensor cannot necessarilycontrol the amount of EGR gas with a good precision when there is adelay in gas transport to the mounting position of the exhaust gassensor or there is a delay in response in the sensor itself.

Therefore, in the present embodiment, by further combining feedbacklearning control based on the air-fuel ratio sensor output with the EGRfeedback control using a combustion parameter, it is possible to controlthe amount of EGR gas with a good response including in transitionaloperation and control the exhaust air-fuel ratio with a good precision.

That is, in the present embodiment, when a predetermined learningcontrol condition (for example, the engine being operated in a steadystate) is satisfied in the state where the feedback control of FIG. 20enables the Δt to be controlled to match with the target value Δt₀, thevalue of the combustion completion period target value Δt₀ is changed alittle at a time so that the exhaust air-fuel detected by the air-fuelratio sensor arranged in the exhaust passage matches with the targetair-fuel ratio determined from the accelerator opening degree Accp andthe engine speed Ne.

For example, when the actual exhaust air-fuel ratio is at the rich sidefrom the target air-fuel ratio, the target value Δt₀ is made to bedecreased by exactly the predetermined amount gt, while when it is atthe lean side from the target air-fuel ratio, the target value Δt₀ ismade to be increased by exactly the predetermined amount gt.

Further, the adjusted target value Δt₀ is used for control of the amountof the EGR gas based on the Δt again so as to adjust the amount of EGRgas so that the actual Δt matches the adjusted target value Δt₀. Whenthe actual Δt and the corrected target value Δt₀ match, it is againjudged if the exhaust air-fuel ratio detected by the air-fuel ratiosensor and the target air-fuel ratio match. When they do not match, thetarget value Δt₀ is again increased or decreased by exactly thepredetermined value gt and the above operation is repeated.

Further, an operation is performed to store the target value Δt₀ whenboth of the exhaust air-fuel ratio and Δt match the target values as thenew target value (learning value) at that accelerator opening degreeAccp and engine speed Ne. By learning correction of the combustionparameter Δt₀ based on the actual air-fuel ratio output in this way, itbecomes possible to accurately control the exhaust air-fuel ratio whilecontrolling the EGR rate with a good response.

Next, another embodiment of the present invention will be explained.

In the embodiments explained above, the PV value was calculated and theΔt found based on PVmax as a combustion parameter was used to controlthe amount of EGR gas. However, as a combustion parameter suitable forcontrol of the amount of EGR gas, it is possible to similarly use, inaddition to PVmax or Δt, any value having a close correlation to one orboth of the ignition delay period and combustion period.

For example, the present embodiment uses as combustion parameters havingclose correlation with the ignition delay period and the combustionperiod the time Atd until the value of PV^(κ) becomes the minimum valuePV^(κ)min and the time Δtc from when the value of PV^(κ) becomes theminimum value PV^(κ)min to when it becomes the maximum value PV^(κ)max.

Here, PV^(κ) is the product of the combustion chamber pressure P at eachcrank angle and the combustion chamber volume V at that crank angle tothe κ power. Further, κ is the specific heat ratio of the air-fuelmixture.

Here, from the equation of state of a gas, in adiabatic change, PV^(κ)becomes constant, but in an actual cylinder compression stroke, heat isradiated through the cylinder wall or through the piston, so in thecylinder compression stroke, PV^(κ) gradually decreases from the startof compression.

On the other hand, if the air-fuel mixture is ignited and combustionstarts, heat of combustion is generated, so the value of PV^(κ) startsto increase. For this reason, the point where the value of PV^(κ)changes from a decrease to increase, that is, the point where PV^(κ)becomes the minimum value PV^(κ)min, is the starting point ofcombustion. Further, similarly, during combustion, the value of PV^(κ)continues to increase, but when the combustion is completed and heat isno longer generated, the value of PV^(κ) again starts to decrease.Therefore, the point where the value of PV^(κ) changes from an increaseto decrease, that is, the point where PV^(κ) becomes the maximum valuePV^(κ)max, is the end point of combustion.

If designating the fuel injection start timing as θinj and the crankangle where the PV^(κ) becomes the minimum value PV^(κ)min as θstart,Δtd=θstart−θinj is the period from the start of fuel injection to thestart of combustion, so becomes equal to the ignition delay period.

Further, if making the crank angle where the PV^(κ) becomes the maximumvalue PV^(κ)max θend, Δtc=θend−θstart becomes equal to the period fromwhen the combustion is started to when it ends, that is, the combustionperiod.

As explained above, the ignition delay period and the combustion periodare both closely correlated with the EGR rate. If the EGR rateincreases, both the ignition delay period and the combustion periodincrease, while if the EGR rate falls, they both decrease.

Therefore, in the present embodiment, either of the ignition delayperiod Δtd or the combustion period Δtc is used to control the EGR rateby a method similar to the case of using the above-mentioned Δt.

That is, in the present embodiment, the value of the ignition delayperiod (or combustion period) at the combustion state giving the optimalEGR rate in advance is set in advance as the target value Δtd₀ (or Δtc₀)for each accelerator opening degree Accp and engine speed Ne. Further,in actual operation, the PV^(κ) is calculated from the combustionchamber pressure and crank angle for every stroke cycle and the crankangle where the value of this PV^(κ) becomes the minimum value (or theminimum value and maximum value) is detected to calculate the Δtd (orΔtc) at actual operation.

Further, the EGR control valve opening degree is feedback controlledbased on the deviation between the Δtd (or Δtc) and the target valueΔtd₀ (or Δtc₀) at the current operating conditions (Accp, Ne).

Note that the specific heat ratio κ can be approximated as a constantvalue and the combustion chamber volume V becomes a function of thecrank angle and can be calculated in advance. Therefore, whencalculating PV^(κ), by calculating the value of V^(κ) for each crankangle in advance and storing it in the ROM of the ECU 20 in the form ofa numerical value table, it is possible to easily calculate the value ofPV^(κ).

Due to this, in the same way as the feedback control using Δt, it ispossible to control the EGR rate accurately with a good response withoutincreasing the processing load of the control circuit.

That is, in the present embodiment, by calculating a combustionparameter corresponding to the combustion timing including at least oneof the ignition delay period or combustion period based on thecombustion chamber pressure detected by the cylinder pressure sensorsand controlling the EGR rate so that the combustion parameter becomes apredetermined target value, it is possible to control the EGR rateaccurately with a good response.

Note that when performing pilot injection injecting a small amount offuel and burning it in the combustion chamber before the main fuelinjection so as to secure good temperature and pressure conditions forburning the main fuel injection fuel, if starting the judgment as towhether the calculated value of the PV^(κ) is the minimum PV^(κ)minafter the start of the main fuel injection, erroneous detection of thepoint of time of start of combustion of the pilot fuel injection fuel asthe point of start of injection of the main fuel injection fuel isprevented.

Next, another embodiment of the present invention will be explained.

In the present embodiment, the control of the fuel injection andinjection timing in the case of multi-fuel injection will be explainedin detail.

As explained above, multi-fuel injection includes pilot injectionperformed before the main fuel injection and after injection etc.performed after the main fuel injection, but both the pilot injectionand after injection may be further divided based on their injectiontimings.

FIG. 23 is a view explaining the different fuel injections forming themulti-fuel injection in the present embodiment.

In FIG. 23, the abscissa shows the crank angle (CA), while the TDC onthe ordinate indicates compression top dead center. Further, theordinate of FIG. 23 shows the injection rates of the different fuelinjections. The areas of the peaks show the relative fuel injectionamounts of the fuel injections. As shown in the figure, with multi-fuelinjection, all or part of early pilot injection, close pilot injection,main injection (main fuel injection), after injection, and postinjection are performed.

Below, the fuel injections other than main injection will be simplyexplained.

(1) Early Pilot Injection

Early pilot injection is pilot injection performed at a timingconsiderably earlier than the main injection (for example, a timingearlier by at least 20 degrees in crank angle (20° CA) than the start ofmain injection). The fuel injected in the early pilot injection forms apremixed air-fuel mixture and ignites by compression, so does notgenerate much NO_(x) or particulate at all. By performing early pilotinjection, it is possible to improve the exhaust properties. Further,early pilot injection raises the temperature and pressure in acombustion chamber and shortens the ignition delay period of the laterexplained close pilot injection or main injection, so can suppress thenoise of combustion or generation of NO_(x) due to the main injection.

Early pilot injection is performed at a point of time when thetemperature and pressure inside the combustion chamber are relativelylow, so when the injection amount is large, the injected fuel reachesthe cylinder wall as a liquid and causes the problem of dilution of thelubrication oil etc. Therefore, when the injection amount is large, theearly pilot injection is performed a plurality of times dividing theamount of injection required into small amounts so as to prevent liquidfuel from reaching the cylinder wall.

(2) Close Pilot Injection

Close pilot injection is pilot injection performed right before the maininjection (for example, within 20° CA from the start of the maininjection). Close pilot injection features less generation ofhydrocarbons compared with early pilot injection and like early pilotinjection shortens the ignition delay period of the main period so cansuppress the noise and generation of NO_(x) of the main injection.

(3) After Injection

After injection is injection started right after the end of maininjection or at a relative short interval from it (for example, within15° CA after the end of main injection).

After injection is designed to increase the temperature, pressure,turbulence, etc. in the combustion chamber again at the end ofcombustion of the fuel of the main injection so as to improve thecombustion and to reduce the injection amount of the main injection.

That is, at the end of combustion of the main injection, the temperatureand pressure in the combustion chamber drop and the turbulence in thecylinder becomes smaller as well, so the fuel becomes harder to burn. Byperforming after injection in this state, an increase in turbulence dueto injection of fuel and an increase in the temperature and fuel due tocombustion of the injected fuel occur, so the atmosphere inside thecombustion chamber is improved in a direction promoting combustion.Further, it is possible to reduce the injection amount of the maininjection by exactly the injection amount of the after injection, so thegeneration of a local excessively region of concentration due to themain injection fuel is suppressed and the drop in the main injectionamount causes the cylinder maximum temperature due to combustion to falland thereby suppresses the generation of NO_(x).

(4) Post Injection

Post injection is fuel injection started after a relative interval afterthe end of the main injection (for example, at least 15° CA after theend of the main injection). The main purpose of the post injection is toraise the exhaust temperature and pressure.

For example, when the temperature of the exhaust purification catalystarranged in the exhaust system is low and the activation temperature isnot reached, so the exhaust gas purification action is not obtained,performing post injection enables the exhaust temperature to be raisedand the catalyst temperature to be raised to the activation temperaturein a short time. Further, by performing the post injection, thetemperature and pressure of the exhaust rise, so in an engine having aturbocharger, it is possible to obtain the effects of an improvement inthe acceleration performance due to the increase in the work of theturbine and the rise in the supercharging pressure and suppression ofsmoke at the time of acceleration.

Further, when using as the exhaust purification catalyst a selectivereduction catalyst purifying the NO_(x) from exhaust using hydrocarboningredients, performing post injection enables hydrocarbons to besupplied to the catalyst and the purification rate of the NO_(x) to beimproved.

As explained above, by performing multi-fuel injection, it is possibleto greatly improve the exhaust properties and noise of a diesel engine,but obtaining this improvement effect requires precise control of theinjection amount and injection timing of each fuel injection in themulti-fuel injection. For example, in the close pilot injection, wherethe greatest precision is required for the injection amount andinjection timing, it is necessary to control a single fuel injectioninjection amount to about 1.5 to 2.5 mm³ and the injection timing toabout ±2° CA or less.

However, as explained above, fuel injectors feature variations betweenindividual injectors due to tolerance, changes in fuel injectioncharacteristics due to period of use, etc., so with normal open loopcontrol, it is not possible to improve the precision of the fuelinjection and the effect of the multi-fuel injection cannot besufficiently obtained.

Further, even if controlling the fuel injection based on the combustionnoise like the system of Japanese Unexamined Patent Publication (Kokai)No. 2001-123871 explained above, while the injection amount of the pilotinjection can be controlled, it is not possible to separately controlthe injection amounts of the different fuel injections. Further, thereis the problem that the injection timing cannot be controlled at all.

In the present embodiment, it is made possible to separately accuratelycontrol the injection among, injection timing, injection pressure, etc.of each fuel injection using the parameters PVY and PV calculated usingthe combustion chamber pressure P detected using the cylinder pressuresensors 29 a to 29 d (hereinafter referred to all together as the“cylinder pressure sensors 29”) and the combustion chamber volume V atthat time.

FIG. 24(A) is a view for explaining the principle of detection of thecombustion timing in the present embodiment. FIG. 24(A) shows the changein the various parameters relating to the combustion in the combustionchamber from the end of the compression stroke of a cylinder to thestart of the expansion stroke with respect to the crank angle θ (crankangle θ=0 indicates compression top dead center). In FIG. 24(A), thecurve P shows the change in the actual combustion chamber pressuredetected by the cylinder pressure sensors 29. Further, the curve Q showsthe heat release rate in a cylinder. As will be understood from curve Q,in the present embodiment, in addition to the main injection, multi-fuelinjection including early pilot injection and after injection isperformed. The peaks Q1, Q2, and Q3 of the heat release rate of FIG.24(A) correspond to the early pilot injection, main injection, and afterinjection.

As will be understood from the curve P of FIG. 24(A), it is possible tojudge the combustion period of each fuel injection at the multi-fuelinjection from the heat release rates Q1, Q2, and Q3, but the combustionperiods of the fuel injections do not clearly appear in the change ofcombustion chamber pressure, so the fuel injections cannot be judgedfrom the curve P.

On the other hand, if calculating the heat release rates, it is possiblein a sense to judge the combustion period of each fuel injection asshown by FIG. 24(A). Further, the heat release rate dQ itself can becalculated from the following equation based on the combustion chamberpressure:dQ/dθ=(κ·P·(dV/dθ)+V(dP/dθ))/(κ−1)

-   -   (where θ shows the crank angle and κ shows the specific heat        ratio of the cylinder air-fuel mixture)

However, calculation of the heat release rate is complicated. Further,it includes many terms including the crank angle θ, so is susceptible toerror of the relatively low detection accuracy crank angle. For thisreason, use of the heat release rate as a control indicator for actualcontrol has problems of the increase in calculation load and error andis not practical.

Therefore, in the present embodiment, the heat release rate is not usedfor detection of the combustion period. The primary rate of change(first derivative) of the value PV^(γ) obtained by multiplying thepressure P detected by the cylinder pressure sensor 29 and the γ powerof the volume V at that time (hereinafter referred to as the “PV^(γ)derivative”) is used. Here, γ is a polytrope exponent.

The polytrope exponent γ may be found in advance by experiments etc.Further, V becomes a function of only θ, so it is also possible tocalculate V^(γ) in advance for the value of each θ. Therefore, thePV^(γ) value can be calculated by simple calculation at each crank angleand the rate of change with respect to θ, that is, the PV^(γ)derivative, can be found by simple calculation of the difference asexplained later.

The curve R of FIG. 24(A) shows the PV^(γ) derivative calculated at eachcrank angle. The PV^(γ) derivative becomes similar in form to the heatrelease rate pattern, so as shown by the curve R, the value of thePV^(γ) derivative becomes mostly zero and becomes a positive value onlyin the part corresponding to the combustion period, so it is possible tojudge the combustion period of each fuel injection extremely clearly.

If approximating the compression due to movement of the piston in thecylinder by the polytrope change of the exponent γ, the pressure P andcombustion chamber volume V have the relationship of PV^(γ)=C (constantvalue). That is, with polytrope change where no combustion occurs and noenergy is given to the gas in the cylinder other than the work due tothe compression, the PV^(γ) value becomes constant at all times. Forthis reason, when no combustion occurs in a combustion chamber, thevalue of the first derivative d(PV^(γ))/dθ of the PV^(γ) value by thecrank angle becomes zero (in actuality, heat is radiated from thecylinder wall, so the value of the first derivative d(PV^(γ))/dθ becomesnegative when no combustion occurs).

On the other hand, if combustion occurs in a cylinder, the gas in thecylinder is given energy (heat) in addition to the work by thecompression, so the change in the gas in the cylinder no longer is apolytrope change. The PV^(γ) value continues increasing while thecombustion is occurring, so the PV^(γ) derivative becomes a positivevalue.

Therefore, the period during which combustion occurs in a combustionchamber can be clearly judged as the period during which the PV^(γ)derivative becomes positive. The same is true, as shown by the curve Rof FIG. 24(A), when a plurality of fuel injections (combustions) Q1, Q2,and Q3 are performed in one stroke cycle of the cylinder.

The present embodiment calculates the PV^(γ) derivative based on thecombustion chamber pressure for each cylinder detected by the cylinderpressure sensor 29 during engine operation and judges the period duringwhich this PV^(γ) derivative is a positive value as the combustionperiod. The “start” at the curve R in FIG. 24(A) shows the combustionstart timing, the “end” shows the combustion end timing, and theinterval between the “start” and “end” shows the combustion period.

Note that the start timing (crank angle) of the combustion period isstrongly correlated with the fuel injection timing. Further, the lengthof the combustion period (length from combustion start timing to endtiming) is strongly correlated with the fuel injection period. Further,the fuel injection period changes due to the injection rate if the fuelinjection amount is constant, while the injection rate changes due tothe fuel injection pressure.

In the present embodiment, the amount of heat release and combustionperiod of each fuel injection giving the optimum combustion state foreach set of engine operating conditions (for example, the combination ofthe engine speed and accelerator opening degree and the type of themulti-fuel injection) are found in advance by experiments etc. and theseoptimum values are stored in the ROM of the ECU 20 as a numerical valuetable using the engine speed and accelerator opening degree for eachfuel injection of the multi-fuel injection (early pilot injection, closepilot injection, main injection, after injection, post injection).

The ECU 20 calculates the PV^(γ) derivative at each crank angle based onthe combustion chamber pressure of each cylinder detected by thecylinder pressure sensor 29 to judge the actual combustion period ofeach injection and feedback controls the fuel injection timing andinjection pressure so that the actual combustion period (start timingand length) becomes the optimum combustion period for the current engineoperating conditions stored in the ROM. Due to this, the fuel injectiontiming and injection pressure of each cylinder are controlled to valuesgiving the optimum combustion period simply and accurately.

Next, the control of the fuel injection amount in the present embodimentwill be explained using FIG. 24(B).

The actual fuel injection amount of each cylinder corresponds to theamount of heat release in the cylinder from the compression stroke tothe expansion stroke of the cylinder. This amount of heat release can becalculated by integrating the heat release rate calculated using theabove equation, but as explained above, calculation using the heatrelease rate dQ is not practical.

Therefore, the actual amount of cylinder heat release is calculatedusing the product PV (hereinafter referred to as the “PV value”) betweenthe combustion chamber pressure P and the combustion chamber volume V atthat time.

The energy of the gas in the combustion chamber is expressed by theproduct PV of the pressure and volume. Therefore, the energy given tothe gas in the combustion chamber per unit crank angle is expressed asd(PV)/dθ.

As explained above, the energy given to the gas in the combustionchamber per unit crank angle becomes the sum of the mechanical energydue to the compression of the piston and the chemical energy generateddue to the combustion.

If converting the equation of state PV=(m/W)RT of gas to a derivativeform, the following is obtained:d(PV)=(m/W)RdT  (1)

-   -   where m is the mass (kg) of the gas in the combustion chamber, w        is the molecular weight of the gas, T is the temperature (K),        and R is a general gas constant (J/mol·K).

Further, equation (1) expresses the amount of change of the energy of agas. This amount of change, as explained above, is expressed as the sumof the mechanical energy d(PV)pist due to compression of the piston andthe chemical energy d(PV)chem generated due to combustion. That is,d(PV)=d(PV)pist+d(PV)chem  (2)

In the above equation (2), the energy d(PV)chem given to a gas bycombustion is expressed as follows using the equation of state of gas ofthe above (1):d(PV)chem=(m/W)RdTchem  (3)

-   -   where, dTchem is the rise in temperature of the gas due to        combustion

If entering equation (3) into equation (2) to find dTchem, the result isdTchem=(W/m·R)(d(PV)−d(PV)pist)  (4)The amount of heat release dQ(J) due to combustion is found as theproduct of the temperature rise dTchem(K), the gas mass m (kg), andconstant volume specific heat (J/mol·K) Cv, so from equation (4),$\begin{matrix}\begin{matrix}{{d\quad Q} = {{m \cdot {Cv} \cdot d}\quad T\quad{chem}}} \\{= {{Cv} \cdot W \cdot {R( {{d({PV})} - {{d({PV})}{pist}}} )}}}\end{matrix} & (5)\end{matrix}$

The amount of heat release ΔQ for each injection can be found byintegrating equation (5) from the combustion start (start) of each fuelinjection to the combustion end (end) since the combustion period foreach injection is known in FIG. 24(A).

That is, $\begin{matrix}\begin{matrix}{{\Delta\quad Q} = {\frac{{CV} \times W}{R}{\int_{start}^{end}{( {{d({PV})} - {d({PV})}_{pist}} )\quad{\mathbb{d}\theta}}}}} \\{= {\frac{{CV} \times W}{R}( {( {({PV})_{end} - ({PV})_{start}} ) -} }} \\ ( {({PV})_{{pist},{end}} - ({PV})_{{pist},{start}}} ) )\end{matrix} & (6)\end{matrix}$

-   -   where, (PV)end and (PV)start are the values at the time of        combustion end and the time of combustion start of the product        of the combustion chamber pressure P detected by the cylinder        pressure sensors and the combustion chamber volume V.

Further, (PV)pist,end and (PV)pist,start are the values at the crankangles corresponding to the time of combustion end (FIG. 24(A), FIG.24(B), end) and the time of combustion start (FIG. 24(A), FIG. 24(B),start) of the product of the combustion chamber pressure in the case ofonly piston compression in the case of no combustion occurring(so-called motoring) and the combustion chamber volume V.

FIG. 24(B) is a view showing the change of the PV value in the case ofFIG. 24(A) and the change of the (PV)pist value. The curve P of FIG.24(B) shows the change of pressure inside the combustion chamber thesame as the curve P of FIG. 24(A), while the curve Q shows the heatrelease rate.

Further, curve S of FIG. 24(B) shows the PV value at the time of changeof pressure of the curve P, while curve T shows the (PV)pist value. The(PV)pist value becomes a constant curve if the engine is determined.

Further, from the above equation (6), for example, the amount of heatrelease ΔQ of the main injection Q1 can be simply found using the PVvalue ((PV)start) and (PV)pist value ((PV)pist,start) at the point A andthe PV value ((PV)end) and (PV)pist value ((PV)pist,end) at the point B.

The amount of heat release ΔQ has a strong correlation with the fuelinjection amount. In the present embodiment, the ideal amount of heatrelease of each injection is found by experiments etc. in advance inaccordance with the engine operating conditions and is stored in the ROMof the ECU 30. Therefore, by feedback correcting the fuel injectionamount so that the actual amount of heat release found from equation (6)matches the ideal amount of heat release stored in the ROM, it becomespossible to control the fuel injection amount to the optimum value.

FIG. 25 is a flow chart showing the actual operation for calculation ofthe combustion period and amount of heat release explained above. Thisoperation is executed every constant crank angle by the ECU 20.

In FIG. 25, at step 401, the current crank angle θ and the combustionchamber pressure P detected by the cylinder pressure sensor 29 are read.Further, at step 403, the current combustion chamber volume V iscalculated based on the crank angle θ. In the present embodiment, therelationship between θ and V is found by calculation in advance and isstored as a one-dimensional numerical value table in the ROM of the ECU20. At step 403, the value of θ read at step 401 is used to find thecombustion chamber volume V from this numerical value table.

Next, at step 405, PV^(γ) is calculated using the pressure P read atstep 401 and the volume V calculated at step 403. γ (polytrope exponent)is found in advance by experiments and is stored in the ROM of the ECU20.

Step 407 shows the operation for calculation of the PV^(γ) derivative.In the present embodiment, the PV^(γ) derivative d (PV^(γ))/dθ iscalculated as the difference between the currently calculated PV^(γ)value (PV^(γ))_(i) and the (PV^(γ))_(i-1) calculated at the time of theprevious execution of this operation.

Next, steps 409 to 417 show detection of the combustion start timing.

At step 409, it is judged if the value of the flag XS has been set to“1”. The flag XS is a flag showing if the detection of the start timingof the combustion period has been completed. XS=1 shows completion ofdetection.

When detection has not been completed at step 409 (X≠1), the routineproceeds to step 411, where it stands by while holding off execution ofthe operation of step 413 on until the PV^(γ) derivative calculated atstep 407 becomes a predetermined value C1 or more. As explained above,the PV^(γ) derivative becomes a value of about zero other than in thecombustion period and becomes a positive value only during thecombustion period. C1 is the judgment value for preventing erroneousdetection due to noise etc. and is set to a positive value as close tozero as possible. When combustion is started, at step 411, the PV^(γ)derivative becomes greater than C1, but when d(PV^(γ))/dθ>C1 firststands at step 411, next at step 413, the value of the flag XS is set to“1”, so from the next time on, steps 411 to 417 are not executed.

Due to this, the combustion start timing is accurately detected. Thatis, when d(PV^(γ))/dθ>C1 first stands at step 411, at steps 415 and 417,the crank angle θ at that time is stored as the crank angle θstart ofthe start of combustion, and the PV value at that time is calculated andstored as the PV value (PV)start at the time of start of combustion.Further, at step 419, the PV value during motoring, that is, the valueof (PV)pist at the time of combustion start, is found from therelationship of the curve T of FIG. 24(B) calculated in advance and isstored as (PV)pist,start.

Steps 419 to 425 show the operation for detection of the end timing ofthe combustion period. The operation for detecting the end timing isperformed only when detection of the start timing of the combustionperiod of steps 411 to 417 is completed and XS is set to “1”.

The operation of steps 419 to 425 is similar to the operation of steps411 to 417, but differs in the point of storing the crank angle whend(PV^(γ))/dθ<C1 at step 419 as the combustion end timing θend and thevalues of (PV) and (PV)pist at that time as (PV)end and (PV)pist,end.Further, after storing the above values, at step 425, the value of theflag XS is set to “0”. Due to this, the steps from step 419 on are notexecuted again until the combustion start timing is detected from steps411 to 417.

Further, after the end of the above operation, at step 427, the aboveequation (6) is used to calculate the amount of heat release ΔQ in thecurrent combustion period.

As explained above, by executing the operation of FIG. 25, thecombustion start timing θstart, combustion end timing θend, and amountof heat release ΔQ are calculated and stored for the plurality of fuelinjections.

Next, the fuel injection control using the combustion start timingθstart, combustion end timing θend, and amount of heat release ΔQcalculated by the above will be explained.

In the present embodiment, the basic values of the fuel injectionamount, injection timing, and fuel injection pressure and the type ofthe injection (type of multi-fuel injection) are set based on apredetermined relationship using the engine speed and acceleratoropening degree by a not shown fuel injection setting operationseparately executed by the ECU 20. Originally speaking, if actuallyinjecting fuel as with these basic values, the combustion state of theengine becomes optimum. However, in practice, due to variation, changes,etc. in the injection characteristics of the fuel injectors, the actualfuel injection does not become like the basic values even if givinginstruction signals corresponding to the basic values to the fuelinjectors.

In the present embodiment, the combustion start timing θstart,combustion end timing θend, and amount of heat release ΔQ are used forfeedback control of the fuel injection so that the actual fuel injectionis performed by the basic values.

FIG. 26 is a flow chart for explaining the routine of the fuel injectioncorrection operation of the present embodiment performed by the ECU 20.

At step 501 of FIG. 26, first the fuel injection desired to be correctedis judged based on the engine operating conditions and the calculatedθstart. That is, it is judged what fuel injection of what type ofmulti-fuel injection (for example, early pilot injection, close pilotinjection, etc.) the fuel injection desired to be corrected is.

Further, at step 503, the target value of the amount of heat release ofthe fuel injection currently desired to be corrected is read from anumerical value table stored in advance in the ROM of the ECU 20 basedon the engine operating conditions (engine speed and accelerator openingdegree).

Further, at step 505, the fuel injection amount of corrected to increaseor decrease so that the amount of heat release ΔQ calculated by theoperation of FIG. 25 explained above matches with the target valueselected at step 503. That is, when the actual amount of heat release ΔQis smaller than the target value, the fuel injection amount is increasedby a predetermined amount, while when it is larger than the targetvalue, it is decreased by a predetermined amount.

Further, at step 507, the target values of the combustion start timingand end timing are similarly read from numerical value tables stored inadvance in the ROM of the ECU 20 in advance based on the engineoperating conditions. At step 509, the fuel injection timing iscorrected so that the actual combustion start timing θstart matches thetarget value. For example, when the combustion start timing is delayedfrom the target value, the fuel injection start timing is advanced,while when it is advanced, it is delayed.

Further, at step 511, the fuel injection pressure is corrected. In thepresent embodiment, the common rail pressure is changed to adjust thefuel injection pressure. That is, at step 511, it is judged if theactual combustion end timing θend is delayed or advanced from the targetvalue in the state with the combustion start timing θstart matched withthe target value. When delayed (when actual combustion period is longerthan target value), the fuel injection pressure is raised by exactly apredetermined amount to cause the end timing of the fuel injection(closing timing of fuel injector) to be advanced by exactly that amountand change the fuel injection period while maintaining the fuelinjection amount constant. Further, conversely, when the actualcombustion end timing is advanced from the target value, the fuelinjection pressure is caused to drop by exactly a predetermined amountto cause the end timing of the fuel injection to be delayed.

By repeating these operations until the values match the target values,the actual fuel injection amount, fuel injection timing, and fuelinjection pressure in the fuel injections of the multi-fuel injectionare feedback corrected so as to become values giving the optimumcombustion in accordance with the operating conditions. As explainedabove, the operation of FIG. 25 enables the combustion timing to befound using a derivative of PV^(γ) able to be calculated by simplecomputation of the difference and the amount of heat release to be foundby simple computation of the PV value. There is no longer a need forcomplicated calculation for detecting the amount of heat release or thecombustion period. Therefore, the increase of the calculation load ofthe ECU 20 is prevented, the amount of heat release and combustionperiod of each injection can be accurately detected simply and reliably,and the injection amount, injection timing, and injection pressure ofeach injection can be accurately feedback controlled.

Further, in the present embodiment, since the injection amounts,injection timings, and injection pressures of the multi-fuel injectionare feedback controlled based on the actual amount of heat release andcombustion period, it becomes possible to accurately correct the fuelinjection characteristics for example even when the variation betweenindividual fuel injection characteristics due to tolerance of the fuelinjectors is relatively large or when the fuel injection characteristicschange along with use. Therefore, even in a common rail type fuelinjection system, variation in characteristics of fuel injectors can betolerated to a certain extent. There is no longer a need to strictlycontrol variation in characteristics of the fuel injectors like in thepast, so the cost of the fuel injectors can be reduced.

As explained above, according to the present invention, by using theoptimum combustion parameter in accordance with the injection mode orcombustion mode for feedback controlling the fuel injection amount,injection timing, and amount of EGR gas, it becomes possible tooptimally control the combustion state of a diesel engine withoutgreatly increasing the processing load of a control circuit.

1. A control system for an internal combustion engine provided with: afuel injector for injecting fuel into an engine combustion chamber, anEGR system for recirculating part of the engine exhaust into the enginecombustion chamber as EGR gas, and a cylinder pressure sensor fordetecting a pressure inside the engine combustion chamber, said controlsystem for an internal combustion engine provided with: combustionparameter calculating means for calculating a combustion parameterexpressing an engine combustion state including at least one of acylinder heat release amount, a combustion start timing, and acombustion period based on a relationship predetermined using thecombustion chamber pressure detected by said cylinder pressure sensorand an engine crank angle and correcting means for correcting at leastone of a fuel injection amount, fuel injection timing, and amount of EGRgas so that the calculated combustion parameter becomes a target valuepredetermined in accordance with the engine operating conditions andselecting as said combustion parameter a combustion parameter giving thesmallest error in accordance with the fuel injection mode or combustionmode of the engine among a plurality of types of combustion parametersexpressing said engine combustion state calculated based on thecombustion chamber pressure and engine crank angle and using saidselected combustion parameter for the correction by said correctingmeans.
 2. A control system of an internal combustion engine as set forthin claim 1, wherein said fuel injection mode includes a fuel injectionmode comprising main fuel injection and multi-fuel injection injectingfuel into the engine combustion chamber before or after or both beforeand after the main fuel injection combined in accordance with need.
 3. Acontrol system of an internal combustion engine as set forth in claim 2,wherein said injection correcting means first corrects an overall fuelinjection amount to the combustion chamber and main fuel injectiontiming and then, when performing multi-fuel injection, corrects the fuelinjection amount or fuel injection timing of said multi-fuel injectionin the state where the main fuel injection has been corrected.
 4. Acontrol system of an internal combustion engine as set forth in claim 2,further provided with sensor calibrating means for correcting error ofsaid cylinder pressure sensor output based on actual combustion chamberpressure detected by said cylinder pressure sensor at a predeterminedcrank angle.
 5. A control system of an internal combustion engine as setforth in claim 2, wherein said engine operating conditions is defined byan engine speed and accelerator opening degree.
 6. A control system ofan internal combustion engine as set forth in claim 5, wherein saidselected combustion parameters are a maximum value of combustion chamberpressure after the start of combustion and a crank angle where thecombustion chamber pressure becomes the maximum.
 7. A control system ofan internal combustion engine as set forth in claim 5, wherein saidselected combustion parameter is a crank angle where a rate of change ofcombustion chamber pressure becomes maximal.
 8. A control system of aninternal combustion engine as set forth in claim 5, wherein saidselected combustion parameter is a crank angle where a second derivativeof combustion chamber pressure becomes maximal.
 9. A control system ofan internal combustion engine as set forth in claim 5, wherein saidselected combustion parameters are a maximum value of a product of thecombustion chamber pressure and combustion chamber actual volume and acrank angle where the product of the combustion chamber pressure andcombustion chamber actual volume becomes maximum.
 10. A control systemof an internal combustion engine as set forth in claim 5, wherein saidselected combustion parameter is a difference ΔPVmax between a maximumvalue of a product of the combustion chamber pressure and combustionchamber actual volume and a product of the combustion chamber pressuredue to only compression in the case of assuming no combustion occurredand combustion chamber actual volume at a crank angle where the productof the combustion chamber pressure and combustion chamber actual volumebecomes maximum.
 11. A control system of an internal combustion engineas set forth in claim 5, wherein said selected combustion parameter is acrank angle where a cylinder heat release rate becomes maximum.
 12. Acontrol system of an internal combustion engine as set forth in claim 5,wherein said selected combustion parameter is an overall amount ofcylinder heat release.
 13. A control system of an internal combustionengine as set forth in claim 5, wherein said selected combustionparameter is a difference between a maximum value of cylinder pressureafter the start of combustion and a cylinder minimum pressure in theinterval after compression top dead center to when combustion is startedin the combustion chamber.
 14. A control system of an internalcombustion engine as set forth in claim 5, wherein said selectedcombustion parameter is a difference between a maximum value of cylinderpressure after the start of combustion and a combustion chamber pressuredue only to compression in the case assuming no combustion occurred atthe crank angle where said cylinder pressure becomes maximum.
 15. Acontrol system of an internal combustion engine as set forth in claim 5,wherein said multi-fuel injection includes pilot fuel injectionperformed before main fuel injection, and said correcting means uses asa combustion parameter the difference between a product of a combustionchamber pressure and combustion chamber actual volume when fuel injectedby main fuel injection is ignited and a product of a combustion chamberpressure due to only compression in the case assuming no combustionoccurred and the combustion chamber actual volume at a crank angle wherefuel injected by the main fuel injection is ignited so as to correct apilot fuel injection instruction.
 16. A control system of an internalcombustion engine as set forth in claim 5, wherein said multi-fuelinjection includes after injection performed after main fuel injection,and said correcting means uses as a combustion parameter the difference(ΔPVmax−ΔPVafter) between a difference ΔPVmax between a maximum value ofa product of a combustion chamber pressure and combustion chamber actualvolume and a product of a combustion chamber pressure when fuel injectedby after injection is ignited and a combustion chamber actual volume anda difference ΔPVafter between a product of a combustion chamber pressurewhen fuel injected by after injection is ignited and the combustionchamber actual volume and the product of a combustion chamber pressuredue to only compression in the case assuming no combustion occurred andthe combustion chamber actual volume at a crank angle where fuelinjected by the after injection is ignited so as to correct an afterinjection instruction.
 17. A control system of an internal combustionengine as set forth in claim 6, wherein said correcting means performssaid correction using as said combustion parameters a maximum value of acombustion chamber pressure after the start of combustion and a crankangle where the combustion chamber pressure becomes maximum only for themain fuel injection.
 18. A control system of an internal combustionengine as set forth in claim 5, wherein said system is further providedwith a throttle valve for throttling the engine intake air amount andwherein said injection correcting means corrects the main fuelinjection, then uses as combustion parameters the difference between thecombustion chamber pressure due to compression at compression top deadcenter and the cylinder minimum pressure in the interval fromcompression top dead center to when combustion is started in thecombustion chamber and the combustion chamber pressure maximum valueafter the start of combustion so as to correct said throttle valveopening degree and main fuel injection timing so that the values ofthese two combustion parameters match their target values.
 19. A controlsystem of an internal combustion engine as set forth in claim 18,wherein said engine can operate switching between a normal combustionmode of injecting fuel after the compression stroke and performingcombustion by a large rate of excess air and a low temperaturecombustion mode of advancing the fuel injection timing from the normalcombustion mode and increasing the amount of EGR gas and wherein saidinjection correcting means corrects said throttle valve opening degreeand main fuel injection timing so that the values of said two combustionparameters match their target values when switching modes between saidnormal combustion mode and low temperature combustion mode.
 20. Acontrol system of an internal combustion engine as set forth in claim 1,wherein said combustion mode of said engine includes modes of differentamounts of supply of EGR gas to the combustion chamber.
 21. A controlsystem of an internal combustion engine as set forth in claim 20,wherein said selected combustion parameter is a time Δt, calculatedbased on a value of a product PV of a combustion chamber pressure Pdetected by a cylinder pressure sensor and a combustion chamber volume Vdetermined from the crank angle θ, from the start of fuel injection fromthe fuel injector to when the value of said PV becomes the maximum valuePVmax and wherein said correcting means adjusts the amount of said EGRgas so that said time Δt becomes a predetermined target value.
 22. Acontrol system of an internal combustion engine as set forth in claim21, wherein further said combustion parameter calculating means furthercalculates a product PVbase between a combustion chamber pressureoccurring due to only compression of the piston in the case of assumingthat combustion did not occur in the combustion chamber and a combustionchamber volume determined from the crank angle and uses a value ofPVbase at the crank angle θpvmax where said PV becomes the maximum valuePVmax to calculate a difference ΔPVmax between PVmax and PVbase andwherein said correcting means further controls the fuel injection amountand fuel injection timing from said fuel injector so that the values ofΔPVmax and said θpvmax become their predetermined target values.
 23. Acontrol system of an internal combustion engine as set forth in claim21, wherein said target value is determined in accordance with an enginespeed and accelerator opening degree.
 24. A control system of aninternal combustion engine as set forth in claim 23, wherein saidinternal combustion engine is a compression ignition engine.
 25. Acontrol system of an internal combustion engine as set forth in claim24, wherein said engine can operate switching between two modes of anormal combustion mode of injecting fuel after the compression strokeand performing combustion by a large rate of excess air and a lowtemperature combustion mode of advancing the fuel, injection timing fromthe normal combustion mode and increasing the amount of EGR gas andcontrols the amount of EGR gas based on the value of said Δt whenoperating in said low temperature mode of the engine.
 26. A controlsystem of an internal combustion engine as set forth in claim 25,wherein said correcting means further makes the fuel injection timingcontinuously change from the injection timing at the normal combustionmode to the target fuel injection timing at the low temperaturecombustion mode over a predetermined transition time when switching fromsaid normal combustion mode to low temperature combustion mode andcontrols the amount of EGR gas based on the value of the Δt calculatedusing the target fuel injection timing at the low temperature combustionmode after switching instead of the actual fuel injection timing duringsaid transition time.
 27. A control system of an internal combustionengine as set forth in claim 20, wherein said selected combustionparameter is a time Δtd, determined based on a value of PV^(κ)calculated from a combustion chamber pressure P detected by saidcylinder pressure sensor, a combustion chamber volume V determined froma crank angle θ, and a specific heat ratio κ of the combustion gas, fromthe start of fuel injection from a fuel injector to when the value ofsaid PV^(κ) becomes the minimum value PV^(κ)min and wherein saidcorrecting means controls the amount of EGR gas so that said Δtd becomesa predetermined target value.
 28. A control system of an internalcombustion engine as set forth in claim 20, wherein said selectedcombustion parameter is a time Δtc, determined based on a value ofPV^(κ) calculated from a combustion chamber pressure P detected by saidcylinder pressure sensor, a combustion chamber volume V determined froma crank angle θ, and a specific heat ratio κ of the combustion gas, fromwhen the value of said PV^(κ) becomes the minimum value PV^(κ)min afterthe start of fuel injection from a fuel injector to when it becomes themaximum value PV^(κ)max and wherein said correcting means adjusts theamount of EGR gas so that said Δtc becomes a predetermined target value.29. A control system of an internal combustion engine as set forth inclaim 27, wherein said fuel injector performs pilot injection forinjecting a small amount of fuel into the combustion chamber before mainfuel injection and wherein said combustion parameter calculating meansstarts detecting the value of said PV^(κ) min after the start of mainfuel injection.
 30. A control system of an internal combustion engine asset forth in claim 1, wherein said selected combustion parameter is arate of change d (PV^(γ))/dθ of the parameter PV^(γ) with respect to thecrank angle θ calculated as a product of a γ power of V and P using acombustion chamber pressure P detected by said cylinder pressure sensor,a combustion chamber volume V determined from the crank angle θ, and apredetermined constant γ and wherein said correcting means detects acombustion period including a combustion start timing and end timing inthe combustion chamber based on said rate of change and corrects atleast one of the fuel injection timing and fuel injection pressure fromsaid fuel injector so that said combustion period matches apredetermined duration.
 31. A control system of an internal combustionengine as set forth in claim 30, wherein further said correcting meanscalculates an amount of cylinder heat release in said combustion periodbased on the value of a parameter PV calculated as a product of saidcombustion chamber pressure P and said combustion chamber volume V atsaid combustion start timing and said combustion end timing and correctsthe fuel injection amount from said fuel injector so that the calculatedamount of cylinder heat release becomes a predetermined value.
 32. Acontrol system of an internal combustion engine as set forth in claim31, wherein said engine performs, in addition to main fuel injection,multi-fuel injection for injecting fuel into the engine combustionchamber before or after or both before and after the main fuel injectionand wherein said correcting means corrects the fuel injection timing orfuel injection pressure based on the value of said d(PV^(γ))/dθ for atleast one fuel injection in the multi-fuel injection and corrects thefuel injection amount based on the value of said PV.